Control system of compression-ignition engine

ABSTRACT

A control system of a compression-ignition engine is provided, which includes an engine configured to cause combustion of mixture gas inside a combustion chamber, a spark plug, and a controller configured to operate the engine. The combustion is performed in a given mode in which, after the spark plug ignites the mixture gas to start combustion, unburned mixture gas combusts by self-ignition. The controller has a heat amount ratio changing module configured to change, according to an engine operating state, a heat amount ratio as an index relating to a ratio of a heat amount generated when the mixture gas combusts by flame propagation with respect to a total heat amount generated when the mixture gas inside the combustion chamber combusts. The controller causes the changing module to increase the heat amount ratio at a high engine speed than at a low engine speed.

TECHNICAL FIELD

The present disclosure relates to a control system of acompression-ignition engine.

BACKGROUND OF THE DISCLOSURE

JP4082292B discloses an engine for combusting a mixture gas inside acombustion chamber by self-ignition within a given operating range wherean engine load and an engine speed are low. The engine combusts themixture gas by spark-ignition within an operating range where the engineload is high and an operating range where the engine speed is higherthan the given operating range.

Incidentally, combustion caused by compression ignition accompanies arelatively loud combustion noise. When the engine speed is high, NVH(Noise Vibration Harshness) of the engine exceeds an allowable value.

SUMMARY OF THE DISCLOSURE

The present disclosure is made in view of the above situations and aimsto perform combustion by compression ignition while suppressing NVH of acompression-ignition engine below an allowable value.

As described later, the present inventors considered a combustion mode(SPCCI (SPark Plug Controlled Compression Ignition) combustion) in whichSI (Spark Ignition) combustion and CI (Compression Ignition) combustionare combined. That is, mixture gas inside a combustion chamber isforcibly ignited to combust through flame propagation, and heatgenerated by this combustion causes unburned mixture gas to combust byself-ignition.

In the CI combustion, the timing of the self-ignition changes greatlydue to a variation in the temperature inside the combustion chamberbefore the compression starts. For example, if the timing of theself-ignition is advanced, the combustion noise increases.

In this regard, the variation in the temperature inside the combustionchamber before the compression starts can be reduced by changing theheat generation amount in the SI combustion. For example, by changingthe ignition timing to change the start timing of the SI combustionaccording to the temperature inside the combustion chamber before thecompression starts, the timing of self-ignition can be controlled.

The combustion by flame propagation causes a relatively small variationin pressure, and thus the combustion noise decreases. Further, the CIcombustion has a shorter combustion period compared to the combustion byflame propagation, which is advantageous in improving fuel efficiency.Therefore, the combustion mode combined the SI combustion and the CIcombustion improves the fuel efficiency while reducing the combustionnoise.

By performing the SPCCI combustion when an engine speed is high, it ispossible to perform the CI combustion while suppressing NVH below theallowable value.

Specifically, according to one aspect of the present disclosure, acontrol system of a compression-ignition engine is provided, whichincludes an engine formed with a combustion chamber and configured tocause combustion of a mixture gas inside the combustion chamber, a sparkplug disposed to be oriented into the combustion chamber and configuredto ignite the mixture gas inside the combustion chamber, and acontroller connected to the spark plug and configured to operate theengine by outputting a control signal to the spark plug.

The combustion is performed in a given mode in which, after the sparkplug ignites the mixture gas to start combustion, unburned mixture gascombusts by self-ignition. The controller has a heat amount ratiochanging module configured to change, according to an operating state ofthe engine, a heat amount ratio as an index relating to a ratio of aheat amount generated when the mixture gas combusts by flame propagationwith respect to a total heat amount generated when the mixture gasinside the combustion chamber combusts. The controller causes the heatamount ratio changing module to increase the heat amount ratio at a highengine speed than at a low engine speed.

Note that the definition of “combustion chamber” here is not limited toa space formed when a piston is at a top dead center on compressionstroke (CTDC) but is broad.

According to this configuration, the controller outputs, once determinedas needed based on the operating state of the engine, the control signalto the spark plug so as to perform the combustion of the given mode inwhich the spark plug ignites the mixture gas inside the combustionchamber at a given timing to start the combustion, and, subsequently,the unburned mixture gas combusts by self-ignition. That is, the SPCCIcombustion described above is performed.

In order to control the SPCCI combustion, the heat amount ratio (SIratio) as the index indicating the characteristic of this combustion isdefined in the controller. Further, the controller includes the heatamount ratio changing module configured to change the heat amount ratioaccording to the operating state of the engine.

When the SPCCI combustion is performed, the heat amount ratio changingmodule controls the SPCCI combustion so that the heat amount ratio at ahigh engine speed becomes higher than at a low engine speed.

As described above, NVH increases at a higher engine speed than at a lowengine speed. Therefore, when the CI combustion which accompanies loudcombustion noise is performed at the high engine speed, similar to whenthe engine speed is low, NVH may exceed the allowable value.

Therefore, when the engine speed is high, by controlling the heat amountratio to be higher, i.e., relatively increasing the ratio of the SIcombustion which accompanies quieter combustion noise in the SPCCIcombustion, so as to relatively reduce the ratio of the CI combustion,NVH is suppressed below the allowable value also at the high enginespeed.

As a more specific configuration, the control system may further includea pressure sensor configured to detect pressure inside the combustionchamber. The controller may set a target heat amount ratio, receive adetection signal of the pressure sensor, calculate the heat amount ratiobased on a pressure waveform caused by the combustion of the mixturegas, and change, when the calculated heat amount ratio is different fromthe target heat amount ratio, the heat amount ratio to approach thetarget heat amount ratio.

Thus, the heat amount ratio is changed according to a difference betweenan actual combustion state inside the combustion chamber based on thedetection signal of the pressure sensor, and a target combustion state.Thus, since the target combustion state of the SPCCI combustioncorresponding to the operating state of the engine is accuratelyachieved, NVH is suppressed below the allowable value.

The control system may further include an internal exhaust gasrecirculation (EGR) system provided to the engine and configured tochange an internal EGR ratio that is a ratio of an amount of internalEGR gas contained within the mixture gas inside the combustion chamber.The controller may output a control signal to the internal EGR system toincrease the internal EGR ratio so as to increase the heat amount ratio.

One method of increasing the heat amount ratio at a high engine speed isincreasing a temperature of the mixture gas inside the combustionchamber (a temperature immediately before the combustion).

At a high engine speed, a combustion cycle is shorter than at a lowengine speed. Since the amount of heat which an inner wall of thecombustion chamber, etc. receives during the combustion accordinglydecreases and the temperature inside the combustion chamber decreases,it is difficult for the mixture gas to receive heat and a time lengthfor the mixture gas to combust becomes short. Therefore, at a highengine speed, the heat amount ratio needs to be increased under such adisadvantageous condition. In this regard, by increasing the temperatureof the mixture gas itself, the SPCCI combustion is stimulated even undersuch a condition.

In the SPCCI combustion, the ignition of the CI combustion occurs afterthe SI combustion. Although increasing the temperature of the mixturegas stimulates both the SI combustion and the CI combustion, thecombustion duration is short when the engine speed is high. Thus, the SIcombustion becomes sharper and is stimulated than the CI combustion,which starts later than the SI combustion. Thus, by increasing thetemperature of the mixture gas, the heat amount ratio is increased alsoat a high engine speed.

Therefore, by increasing the internal EGR ratio so as to increase theratio of the internal EGR gas contained within the mixture gas(high-temperature burned gas), the temperature of the mixture gasincreases, and thus, the heat amount ratio increases. As a result, NVHis suppressed below the allowable value also at a high engine speed.

Further, the control system may further include an external EGR systemprovided to the engine and configured to change an external EGR ratiothat is a ratio of an amount of external EGR gas contained within themixture gas inside the combustion chamber. The controller may output acontrol signal to the external EGR system to reduce the external EGRratio so as to increase the heat amount ratio.

While the engine is operating, also by reducing the external EGR ratioso as to reduce the ratio of the external EGR gas contained within themixture gas (cooled low-temperature burned gas), the temperature of themixture gas increases, and thus, the heat amount ratio increases. As aresult, NVH is suppressed below the allowable value also at a highengine speed.

By increasing the internal EGR ratio and reducing the external EGRratio, the temperature of the mixture gas before the combustion is moreeffectively increased, and thus, the heat amount ratio is stablyincreased.

A swirl flow may be formed in the combustion mode when the combustion isperformed in the given mode.

If the swirl flow is formed in the combustion chamber, the flow of themixture gas inside the combustion chamber is stimulated and also themixture gas in the combustion chamber is stratified. Thus, hazardoussubstances, such as NO_(x) and soot, are prevented from being generated,and the mixture gas by which the SI combustion in the SPCCI combustionis stably performed is formed around the spark plug at an ignitiontiming. Therefore, the degree of freedom of controlling the heat amountratio is increased and a more stable SPCCI combustion is achieved.

The controller may output the control signal to the spark plug toadvance an ignition timing so as to increase the heat amount ratio.

When the ignition timing is advanced, the SI combustion starts early.Thus, the SI combustion is stimulated even in a short period of time ata high engine speed, and the heat amount ratio is increased. This methodmay be combined with the methods described above, and thereby, the heatamount ratio at a high engine speed is more effectively increased.

The internal EGR gas may be introduced into the combustion chamber byoverlapping an open period of an intake valve with an open period of anexhaust valve.

In the introduction of the internal EGR gas by setting a so-calledpositive overlap period in which the open period of the intake valve andthe open period of the exhaust valve overlap, the high-temperatureburned gas generated in the combustion chamber once flows from thecombustion chamber to an intake passage. Then, the burned gas flowedinto the intake passage is introduced again into the combustion chamber.By flowing to the intake passage where the temperature is low, thehigh-temperature burned gas to be introduced into the combustionchamber, i.e., the internal EGR gas, is cooled.

Therefore, the introduction of the internal EGR gas by setting thepositive overlap period lowers the temperature of the combustion chambermore than when the internal EGR gas is introduced by setting a negativeoverlap period in which the high-temperature burned gas is confined asit is in the combustion chamber. By this temperature decrease, itbecomes possible to introduce a larger amount of the internal EGR gas,and thus, the CI combustion by self-ignition becomes slower. As aresult, the degree of freedom of controlling the heat amount ratio isincreased and the stable SPCCI combustion is achieved.

A swirl ratio of the swirl flow may be below 4.

In this case, for example, the control system may further include aswirl control valve provided to the engine and configured to change astrength of the swirl flow by controlling an opening by the controller.The strength of the swirl flow may increase as the opening of the swirlcontrol valve is reduced. The controller may control the swirl controlvalve to have a given narrow opening without fully closing the swirlcontrol valve.

Although forming the swirl flow in the combustion chamber stimulates theflow of the mixture gas inside the combustion chamber and stratify themixture gas in the combustion chamber as described above, when the swirlflow is excessively strong, the stratification of the mixture gas may beunstable depending on the combustion condition. When the swirl ratio isbelow 4, the flow of the mixture gas inside the combustion chamberbecomes suitable, and thus, the stable stratification of the mixture gasis achieved in a wide operating range of the engine.

Specifically, according to another aspect of the present disclosure, acontrol system of a compression-ignition engine is provided, whichincludes an engine formed with a combustion chamber and configured tocause combustion of mixture gas inside the combustion chamber, a sparkplug disposed to be oriented into the combustion chamber and configuredto ignite the mixture gas inside the combustion chamber, an EGR systemprovided to the engine and configured to change an EGR ratio that is aratio of an amount of EGR gas contained within the mixture gas insidethe combustion chamber, and a controller connected to the spark plug andthe EGR system and configured to operate the engine by outputting acontrol signal to the spark plug and the EGR system, respectively.

The controller outputs the control signal to the spark plug at a givenignition timing so as to perform the combustion in a given mode inwhich, after the spark plug ignites the mixture gas to start combustion,unburned mixture gas inside the combustion chamber combusts byself-ignition. The controller outputs the control signal to the EGRsystem so as to increase the EGR ratio higher at a high engine speedthan at a low engine speed.

According to this configuration, the controlled outputs, once determinedas needed based on the operating state of the engine, the control signalto the spark plug so as to perform the combustion of the given mode inwhich the spark plug ignites the mixture gas inside the combustionchamber at a given timing to start the combustion, and, subsequently,the unburned mixture gas combusts by self-ignition. That is, the SPCCIcombustion described above is performed.

When the SPCCI combustion is performed, the controller outputs thecontrol signal to the EGR system provided to the engine. The EGR systemchanges the EGR ratio (a ratio of an amount of EGR gas contained withinthe mixture gas inside the combustion chamber) and increases the EGRratio to be higher at a high engine speed than at a low engine speedaccording to the control signal outputted from the controller.

As described above, NVH increases at a high engine speed than at a lowengine speed. Therefore, when the CI combustion which accompanies loudcombustion noise is performed at the high engine speed, similar to whenthe engine speed is low, NVH may exceed the allowable value. Therefore,in order to perform the combustion by the self-ignition at a high enginespeed, it is required to reduce a ratio of the CI combustion in theSPCCI combustion while increasing a ratio of the SI combustion.

At a high engine speed, a combustion cycle is shorter than at a lowengine speed. Since the amount of heat that an inner wall of thecombustion chamber, etc. receives during the combustion accordinglydecreases and the temperature inside the combustion chamber decreases,it is difficult for the mixture gas to receive heat and a time lengthfor the mixture gas to combust becomes short. Therefore, at a highengine speed, the SPCCI combustion needs to be stimulated to increasethe ratio of the SI combustion under such a disadvantageous condition.In this regard, by increasing the temperature of the mixture gas itself(the temperature immediately before the combustion), the SPCCIcombustion is stimulated even under such a condition.

In the SPCCI combustion, the ignition of the CI combustion occurs afterthe SI combustion. Although increasing the temperature of the mixturegas stimulates both the SI combustion and the CI combustion, thecombustion duration is short when the engine speed is high. Thus, the SIcombustion becomes sharper and is stimulated more than the CIcombustion, which starts later than the SI combustion. Thus, byincreasing the temperature of the mixture gas, the ratio of the SIcombustion is increased at a high engine speed.

Therefore, while the engine is operating, by increasing the EGR ratio soas to increase the ratio of the EGR gas contained within the mixture gas(high-temperature burned gas), the temperature of the mixture gasincreases, and thus, the ratio of the SI combustion in the SPCCIcombustion increases. As a result, NVH is suppressed below the allowablevalue also at a high engine speed. Note that the EGR gas may be eitherone of the internal EGR gas and the external EGR gas as long as it has ahigh temperature.

When the engine speed exceeds a given limitation starting speed, thecontroller may output the control signal to the EGR system so as tolimit the increase in the EGR ratio.

Although increasing the EGR ratio raises the temperature of the mixturegas, which suppresses NVH below the allowable value as described above,if the temperature of the mixture gas rises excessively, the CIcombustion becomes sharp and the combustion noise increases, which maycause NVH above the allowable value. Therefore, the controller limitsthe increase in the EGR ratio before such a situation may occur so as toprevent NVH from exceeding the allowable value.

In this case, the limitation starting speed may be set to be lower asthe engine load increases.

As the engine load increases, the temperature of the mixture gas easilybecomes excessively high. Therefore, by setting the limitation startingspeed at which the limitation of the increase in the EGR ratio starts,lower as the engine load increases, it is more surely prevented that NVHexceeds the allowable value.

The control system may further include an intake flow control deviceattached to the engine and configured to change a flow of intake airintroduced into the combustion chamber. The controller may output acontrol signal to the intake flow control device to increase the flow ofthe intake air within an engine speed range exceeding the limitationstarting speed.

Above the limitation starting speed, since the ratio of the SIcombustion does not increase, the effect of reducing the combustionnoise is limited and the effect may not be obtained sufficiently.Therefore, in accordance with the above configuration, the effect whichis limited by changing the EGR ratio is compensated. That is, bystrengthening the intake air flow, a vaporization of injected fuel isstimulated and the SI combustion is performed in a state where the flowinside the combustion chamber is strong. Thus, the SI combustion becomessharp and the SI combustion is stimulated compared to the CI combustionin which the ignition is performed later. Therefore, the ratio of the SIcombustion in the SPCCI combustion increases and it is more surelyprevented that NVH exceeds the allowable value.

Further, the controller may output a control signal to the spark plug sothat an ignition timing advances within an engine speed range exceedingthe limitation starting speed.

When the ignition timing is advanced, the SI combustion starts early.Thus, the SI combustion becomes sharp even in a short period of time ata high engine speed, which allows the ratio of the SI combustion in theSPCCI combustion to be increased. Therefore, also by advancing theignition timing, the effect which is limited by changing the EGR ratiois compensated and NVH is suppressed below the allowable value also at ahigh engine speed.

The EGR gas may be internal EGR gas.

Since the internal EGR gas is introduced into the combustion chamberdirectly, the high temperature is secured more easily than the externalEGR gas introduced into the combustion chamber indirectly. Therefore,the temperature of the mixture gas is easily increased, and the ratio ofthe SI combustion in the SPCCI combustion is easily increased.

In this case, the internal EGR gas may be introduced into the combustionchamber by providing a negative overlap period.

Alternatively, the internal EGR gas may be introduced into thecombustion chamber by overlapping an open period of an intake valve withan open period of an exhaust valve (so-called a positive overlapperiod).

In the introduction of the internal EGR gas by setting a negativeoverlap period (NVO setting), burned gas residing in the combustionchamber is used as the internal EGR gas as it is, the temperature of themixture gas is easily increased, and the ratio of the SI combustion inthe SPCCI combustion is easily increased.

On the other hand, in the introduction of the internal EGR gas bysetting the positive overlap period (PVO setting), the high-temperatureburned gas generated in the combustion chamber once flows from thecombustion chamber to an intake passage. Then, the burned gas flowedinto the intake passage is introduced again into the combustion chamber.By flowing to the intake passage where the temperature is low, thehigh-temperature burned gas introduced into the combustion chamber,i.e., the internal EGR gas, is cooled.

Therefore, the introduction of the internal EGR gas by the PVO settinglowers the temperature of the combustion chamber more than the NVOsetting. Thus, the CI combustion by self-ignition in the SPCCIcombustion becomes slower and the increase of the combustion noise dueto excessive CI combustion is prevented.

Further in this case, the EGR ratio may continuously increase withoutlimitation until the engine speed comes close to a highest speed withinan operating range of the engine in which the combustion of the givenmode is performed.

As described above, if the temperature of the mixture gas is increasedexcessively, since NVH may exceed the allowable value, the increase inthe EGR ratio needs to be limited. In this regard, in the introductionof the internal EGR gas by the NVO setting, since the temperature in thecombustion chamber easily becomes excessively high, the necessity oflimiting the increase in the EGR ratio is high. On the other hand, inthe introduction of the internal EGR gas by the PVO setting, thetemperature in the combustion chamber is prevented from increasingexcessively high. Thus, in the introduction of the internal EGR gas bythe PVO setting, the necessity of limiting the increase in the EGR ratiois low, and the EGR ratio may be continuously high until the enginespeed comes close to the highest speed.

The control system may further include a boosting system attached to theengine and configured to boost gas to be introduced into the combustionchamber. The controller may output a control signal to the boostingsystem so as to perform boosting within a first range in which theengine load is higher than a given load, and not to perform the boostingwithin a second range in which the engine load is below the given load.Within the second range, the controller may output the control signal tothe EGR system so as to increase the EGR ratio higher at a high enginespeed than at a low engine speed.

With such an engine, within the first range in which the boosting isperformed, even when the EGR gas is introduced, it is scavenged by theboosting pressure. Therefore, within the first range, it is difficult tochange the EGR ratio, a state function inside the combustion chamberbecomes unstable and may cause a degradation of fuel efficiency, etc.Therefore, within the second range in which the boosting is notperformed, the controller outputs the control signal to the EGR systemso that the EGR ratio becomes higher. Thus, NVH is suppressed below theallowable value also at a high engine speed.

A swirl flow may be formed in the combustion chamber when the combustionis performed in the given mode.

If the swirl flow is formed in the combustion chamber, the flow of themixture gas inside the combustion chamber is stimulated and also themixture gas in the combustion chamber is stratified. Thus, hazardoussubstances, such as NO_(x) and soot, are prevented from being generated,and the mixture gas by which the SI combustion in the SPCCI combustionis stably performed is unevenly formed around the spark plug at anignition timing. Therefore, the stable SPCCI combustion is achieved andNVH is suppressed below the allowable value also at a high engine speed.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a diagram illustrating a configuration of an engine.

FIG. 2 is a diagram illustrating a structure of a combustion chamber, inwhich the upper part is a plan view of the combustion chamber and thelower part is a II-II cross-sectional view.

FIG. 3 is a plan view illustrating structures of the combustion chamberand an intake system.

FIG. 4 is a block diagram illustrating a configuration of a controldevice of the engine.

FIG. 5 is a diagram illustrating a rig test device for measuring a swirlratio.

FIG. 6 is a chart illustrating a relationship between an opening ratioof a secondary passage and the swirl ratio.

FIG. 7A is a chart illustrating a first operating range map.

FIG. 7B is a chart illustrating a second operating range map.

FIG. 7C is a chart illustrating a third operating range map.

FIG. 8 shows charts conceptually illustrating a change in heatgeneration rate in SPCCI combustion in which SI combustion and CIcombustion are combined.

FIG. 9 is a chart illustrating a change in an SI ratio, a change in astate function inside a combustion chamber, a change in an overlapperiod between an intake valve and an exhaust valve, and changes in aninjection timing and ignition timing of fuel, with respect to an engineload corresponding to the first operating range map.

FIG. 10 shows charts in which the upper part illustrates a change in acombustion waveform according to an increase in the engine load innon-boosted SPCCI combustion, and the lower part illustrates a change ina combustion waveform according to an increase in the engine load inboosted SPCCI combustion.

FIG. 11 is a chart illustrating one example of a relationship between anengine speed and the SI ratio within an operating range in which theSPCCI combustion of the engine corresponding to the first operatingrange map is performed.

FIGS. 12A and 12B show charts illustrating one example of a relationshipbetween the engine speed and an internal EGR ratio within the operatingrange in which the SPCCI combustion of the engine corresponding to thefirst operating range map is performed, in which FIG. 12A illustratesthe relationship when the internal EGR gas is introduced by setting anegative overlap period, and FIG. 12B illustrates the relationship whenthe internal EGR gas is introduced by setting a positive overlap period.

FIGS. 13A and 13B show charts illustrating one example of a relationshipbetween the engine speed and an external EGR ratio within the operatingrange in which the SPCCI combustion of the engine corresponding to thefirst operating range map is performed, in which FIG. 13A illustratesthe relationship when a swirl flow is scarcely formed in the combustionchamber, and FIG. 13B illustrates the relationship when a swirl flowwith a given strength is formed in the combustion chamber.

FIGS. 14A and 14B show charts illustrating one example of a relationshipbetween the engine speed and the opening of a swirl control valve withinthe operating range in which the SPCCI combustion of the enginecorresponding to the first operating range map is performed, in whichFIG. 14A illustrates the relationship in the engine corresponding to thefirst operating range map, and FIG. 14B illustrates the relationship inthe engine corresponding to a second or third operating range map.

FIG. 15 is a chart illustrating one example of a relationship betweenthe engine speed and the ignition timing within the operating range inwhich the SPCCI combustion of the engine corresponding to the firstoperating range map is performed.

FIG. 16 is a flowchart illustrating a flow of a control of the engineexecuted by an ECU.

FIG. 17 is a diagram illustrating a control concept regarding a changeof the SI ratio.

FIG. 18 is a diagram illustrating fuel injection timings, ignitiontimings, and combustion waveforms in respective operating states on thethird operating range map.

FIG. 19 is a brief chart of the third operating range map illustrated inFIG. 7C, illustrating respective operating states.

FIG. 20 is a diagram illustrating combustion waveforms in the respectiveoperating states illustrated in FIG. 19.

DETAILED DESCRIPTION OF THE DISCLOSURE

Hereinafter, embodiments of a control system of a compression-ignitionengine (hereinafter, may simply be referred to as “engine 1”) aredescribed in detail with reference to the accompanying drawings. Thefollowing description gives one example of the control system of theengine 1. FIG. 1 is a diagram illustrating a configuration of the engine1. FIG. 2 is a cross-sectional view illustrating a structure of acombustion chamber. FIG. 3 is a plan view illustrating structures of thecombustion chamber and an intake system. Note that in FIG. 1, an intakeside is on the left side and an exhaust side is on the right side of thedrawing sheet. Further in FIGS. 2 and 3, the intake side is on the rightside and the exhaust side is on the left side of the drawing sheets.FIG. 4 is a block diagram illustrating a configuration of the controlsystem of the engine 1.

Further, the definition of “EGR gas” used below includes “burned gas(exhaust gas) that remains in the combustion chamber and/or is suckedinto the combustion chamber again.” Similarly, the definition of“internal EGR gas” includes “burned gas (exhaust gas) that remains inthe combustion chamber and/or is sucked directly into the combustionchamber again without flowing through a passage outside the engine,” andthe definition of “external EGR gas” includes “burned gas (exhaust gas)indirectly sucked into the combustion chamber again via the passageoutside the engine, such as an exhaust passage and an intake passage.”Further, “EGR ratio” is equivalent to “a ratio of an amount of EGR gascontained within a mixture gas (total gas) in the combustion chamber.”Moreover, “internal EGR ratio” is equivalent to “a ratio of an amount ofthe internal EGR gas contained within the mixture gas (total gas) in thecombustion chamber” and “external EGR ratio” is equivalent to “a ratioof an amount of the external EGR gas contained within the mixture gas(total gas) in the combustion chamber.” Details of “heat amount ratio(SI ratio)” are described later.

<SPCCI Combustion>

The engine 1 performs combustion in a mode in which SI combustion and CIcombustion are combined.

The SI combustion is combustion accompanying flame propagation whichstarts by forcibly igniting mixture gas inside a combustion chamber. TheCI combustion is combustion which starts by the mixture gas inside thecombustion chamber igniting by being compressed. In the combustion modecombined the SI combustion and the CI combustion, a spark plug forciblyignites the mixture gas inside the combustion chamber to combust itthrough flame propagation, and heat generated by this combustion andpressure increase thereby raise the temperature inside the combustionchamber, which leads to combustion of unburned mixture gas byself-ignition.

In this regard, the variation in the temperature inside the combustionchamber before the compression starts can be reduced by adjusting theheat generation amount in the SI combustion. For example, by controllingthe ignition timing to adjust the start timing of the SI combustionaccording to the temperature inside the combustion chamber before thecompression starts, the unburned mixture gas can self-ignite at a targettiming.

Hereinafter, the combustion mode in which the SI combustion and the CIcombustion are combined so that the CI combustion is controlled usingthe SI combustion is referred to as SPCCI (Spark Controlled CompressionIgnition) combustion.

<Configuration of Engine 1>

The engine 1 is mounted on a four-wheel automobile. The automobiletravels by the operation of the engine 1. Fuel of the engine 1 isgasoline in this embodiment. The gasoline may contain bioethanol, etc.The fuel of the engine 1 may be any kind of fuel as long as it is liquidfuel containing at least gasoline.

The engine 1 includes a cylinder block 12 and a cylinder head 13 placedon the cylinder block 12. The cylinder block 12 is formed therein with aplurality of cylinders 11. In FIGS. 1 and 2, only one cylinder 11 isillustrated. The engine 1 is a multi-cylinder engine. The engine 1 alsoincludes a piston 3, an injector 6, a spark plug 25, an intake valve 21,and an exhaust valve 22.

(Piston 3)

The piston 3 is reciprocatably inserted in each cylinder 11. The piston3 is coupled to a crankshaft 15 via a connecting rod 14. The piston 3defines a combustion chamber 17 together with the cylinder 11 and acylinder head 13. Note that the definition of “combustion chamber” isnot limited to a space formed when the piston 3 is at a top dead centeron compression stroke (CTDC) and may be broad. That is, “combustionchamber” may mean any space formed by the piston 3, the cylinder 11, andthe cylinder head 13 regardless of the position of the piston 3.

As illustrated in FIG. 2, a lower surface of the cylinder head 13, thatis, a ceiling surface of the combustion chamber 17, is formed by aninclined surface 1311 and an inclined surface 1312. The inclined surface1311 inclines toward an axis X2 (an axis passing through the center ofinjection of the injector 6) from the intake side. The inclined surface1312 inclines upwardly toward the axis X2 from the exhaust side. Theceiling surface of the combustion chamber 17 has a so-called pent-roofshape.

An upper surface of the piston 3 bulges toward the ceiling surface ofthe combustion chamber 17 (or may be flat). A cavity 31 is formed in theupper surface of the piston 3. The cavity 31 is formed by denting theupper surface of the piston 3. The cavity 31 has a shallow plate shape.The cavity 31 faces the injector 6 when the piston 3 is located nearCTDC.

The center of the cavity 31 is offset from an axis X1 (an axis passingthrough the radial center of the cylinder 11) to the exhaust side andcoincides with the axis X2. The cavity 31 has a convex section 311. Theconvex section 311 is formed on the axis X2. The convex section 311 hasa substantially conical shape. The convex section 311 extends upwardlytoward the ceiling surface of the cylinder 11 from the bottom of thecavity 31.

The cavity 31 has a dented section 312 formed to surround the convexsection 311 entirely. The cavity 31 has a symmetric shape with respectto the axis X2. A squish area 171 is formed outside a section of thecavity 31 within the combustion chamber 17.

A circumferential side face of the dented section 312 extends from abottom surface of the cavity 31 to an opening surface of the cavity 31,inclined with respect to the axis X2. An inner diameter of the cavity 31at the dented section 312 gradually increases from the bottom surface ofthe cavity 31 to the opening surface of the cavity 31.

Note that the shape of the combustion chamber 17 is not limited to thatillustrated in FIG. 2. The shapes of the cavity 31, the upper surface ofthe piston 3, the ceiling surface of the combustion chamber 17, etc. aresuitably changeable.

For example, the depth of the dented section 312 may be shallower on theouter circumferential side. In this case, an amount of EGR gas aroundthe spark plug 25 decreases, and flame propagation of SI combustion inthe SPCCI combustion becomes favorable.

As indicated by a virtual line L1 of FIG. 2, the cavity 31 may also havean asymmetric shape with respect to the axis X2. That is, the dentedsection 312 is formed radially larger and deeper on the exhaust sidethan the intake side. In this case, a fuel concentration around thespark plug 25 at the time of ignition increases, and therefore,ignitability of the SI combustion in the SPCCI combustion improves.

As indicated by a virtual line L2 of FIG. 2, the cavity 31 may not havethe convex section 311. That is, the cavity 31 has a spherical shapegradually becoming shallower radially outwardly from the center. In thiscase, since it becomes more difficult for the piston 3 to come intocontact with the intake valve 21 and the exhaust valve 22, the degree offreedom of controlling the opening and closing of the intake valve 21and the exhaust valve 22 increases. If a swirl flow is formed in thecombustion chamber 17 in this case, the flow stabilizes, and therefore,stratification of the mixture gas becomes easy.

The geometric compression ratio of the engine 1 is set to be between 13and 30. As described above, the SPCCI combustion controls the CIcombustion by utilizing the heat generated by the SI combustion and thepressure increase. In this engine 1, it is unnecessary to raise thetemperature of the combustion chamber 17 when the piston 3 reaches CTDCfor the mixture gas to self-ignite (i.e., the compression endtemperature).

That is, although the engine 1 performs the CI combustion, its geometriccompression ratio may be set relatively low. Lowering the geometriccompression ratio is advantageous in reducing a cooling loss and amechanical loss. For example, the geometric compression ratio may be setto 14:1 to 17:1 in regular specifications (the octane number of the fuelis about 91) and to 15:1 to 18:1 in high-octane specifications (theoctane number of the fuel is about 96).

(Intake Valve 21 and Exhaust Valve 22)

The cylinder head 13 is formed with an intake port 18 for each cylinder11. As illustrated in FIG. 3, the intake port 18 includes two intakeports of a first intake port 181 and a second intake port 182. The firstintake port 181 and the second intake port 182 are arranged in axialdirections of the crankshaft 15, i.e., front-and-rear directions of theengine 1. The intake port 18 communicates with the combustion chamber17. Although not illustrated in detail, the intake port 18 is aso-called tumble port. That is, the intake port 18 has such a shape thata tumble flow is formed in the combustion chamber 17.

The intake valve 21 is disposed in the intake port 18. The intake valve21 opens and closes the intake port 18 to and from the combustionchamber 17. The intake valve 21 is opened and closed by a valveoperating mechanism at a given timing. This valve operating mechanismmay be a variable valve mechanism which makes a valve timing and/orvalve lift variable. In this configuration example, as illustrated inFIG. 4, the variable valve mechanism has an intake electrically-operatedS-VT (Sequential-Valve Timing) 23. The intake electrically-operated S-VT23 is continuously variable of a rotational phase of an intake camshaftwithin a given angular range. Thus, the open and close timings of theintake valve 21 continuously change. Note that the intake valveoperating mechanism may have a hydraulically-operated S-VT instead ofthe electrically-operated S-VT.

The cylinder head 13 is also formed with an exhaust port 19 for eachcylinder 11. As illustrated in FIG. 3, the exhaust port 19 also includestwo exhaust ports of a first exhaust port 191 and a second exhaust port192. The first exhaust port 191 and the second exhaust port 192 arearranged in the front-and-rear directions of the engine 1. The exhaustport 19 communicates with the combustion chamber 17.

The exhaust valve 22 is disposed in the exhaust port 19. The exhaustvalve 22 opens and closes the exhaust port 19 to and from the combustionchamber 17. The exhaust valve 22 is opened and closed by a valveoperating mechanism at a given timing. This valve operating mechanismmay be a variable valve mechanism which makes a valve timing and/orvalve lift variable. In this configuration example, as illustrated inFIG. 4, the variable valve mechanism has an exhaustelectrically-operated S-VT 24. The exhaust electrically-operated S-VT 24is continuously variable of a rotational phase of an exhaust camshaftwithin a given angular range. Thus, the open and close timings of theexhaust valve 22 continuously change. Note that the exhaust valveoperating mechanism may have a hydraulically-operated S-VT instead ofthe electrically-operated S-VT.

Although is described later in detail, the engine 1 adjusts the lengthof an overlap period of the open timing of the intake valve 21 and theclose timing of the exhaust valve 22 by the intake electrically-operatedS-VT 23 and the exhaust electrically-operated S-VT 24. Thus, residualgas in the combustion chamber 17 is scavenged. Further, by adjusting thelength of the overlap period, internal EGR gas is introduced into thecombustion chamber 17 or is confined in the combustion chamber 17. Inthis configuration example, the intake electrically-operated S-VT 23 andthe exhaust electrically-operated S-VT 24 constitute an internal EGRsystem as one of state function setting devices. Note that the internalEGR system is not necessarily constituted by the S-VT.

(Injector 6)

The injector 6 is attached to the cylinder head 13 for each cylinder 11.The injector 6 injects the fuel directly into the combustion chamber 17.The injector 6 is disposed in a valley portion of the pent roof wherethe inclined surface 1311 on the intake side and the inclined surface1312 on the exhaust side intersect. As illustrated in FIG. 2, theinjector 6 is disposed so that its injection axis extends along the axisX2. The injection axis of the injector 6 coincides with the position ofthe convex section 311 of the cavity 31. The injector 6 is orientedtoward the cavity 31. Note that the injection axis of the injector 6 maycoincide with the center axis X1 of the cylinder 11. Also in this case,it is desirable that the injection axis of the injector 6 coincides withthe position of the convex section 311 of the cavity 31.

Although is not illustrated in detail, the injector 6 is constructed bya multi-port fuel injector having a plurality of nozzle ports. Asindicated by two-dotted chain lines in FIG. 2, the injector 6 injectsthe fuel so that the fuel spray radially spreads downward from an uppercenter section of the combustion chamber 17 where the injection centeris located. In this configuration example, the injector 6 has ten nozzleports, and the nozzle ports are arranged at an even angular interval inthe circumferential direction.

As illustrated in the upper part of FIG. 2, the axes of the nozzle portsdo not circumferentially overlap with the spark plug 25 described later.That is, the spark plug 25 is sandwiched between the axes of twoadjacent nozzle ports. Thus, the fuel spray injected from the injector 6is prevented from directly hitting the spark plug 25 and getting anelectrode wet.

As described later, the injector 6 may inject the fuel at the timingwhen the piston 3 is positioned near CTDC. In this case, when theinjector 6 injects the fuel, the fuel spray flows downward along theconvex section 311 of the cavity 31 while mixing with fresh air, andflows along the bottom surface and the circumferential surface of theconcave portion 312 to spread radially outward from the center of thecombustion chamber 17. Then, the mixture gas reaches the opening of thecavity 31, flows along the inclined surface 1311 on the intake side andthe inclined surface 1312 on the exhaust side, and further flows fromthe outer circumferential side toward the center of the combustionchamber 17. Note that the injector 6 is not limited to the multi-portinjector. The injector 6 may adopt an outward-opening valve injector.

A fuel supply system 61 is connected to the injector 6. The fuel supplysystem 61 includes a fuel tank 63 configured to store the fuel and afuel supply path 62 connecting the fuel tank 63 with the injector 6. Afuel pump 65 and a common rail 64 are provided in the fuel supply path62. The fuel pump 65 pumps the fuel to the common rail 64. In thisembodiment, the fuel pump 65 is a plunger pump which is driven by thecrankshaft 15.

The common rail 64 stores the fuel pumped from the fuel pump 65 at highfuel pressure. When the injector 6 opens, the fuel stored in the commonrail 64 is injected into the combustion chamber 17 from the nozzle portsof the injector 6. The fuel supply system 61 is suppliable of the fuelat a high pressure of 30 MPa or higher to the injector 6. A highest fuelpressure of the fuel supply system 61 may be, for example, about 120MPa. The pressure of the fuel supplied to the injector 6 may be changedaccording to an operating state of the engine 1. Note that the structureof the fuel supply system 61 is not limited to the above structure.

(Spark Plug 25)

The spark plug 25 is attached to the cylinder head 13 for each cylinder11. The spark plug 25 forcibly ignites the mixture gas in the combustionchamber 17. In this configuration example, the spark plug 25 is disposedat the intake side of the center axis X1 of the cylinder 11. The sparkplug 25 is located between the two intake ports 18. The spark plug 25 isattached to the cylinder head 13 to extend downwardly, toward the centerof the combustion chamber 17 in a tilted posture with respect toup-and-down directions of the cylinder head 13.

As illustrated in FIG. 2, the electrode of the spark plug 25 is locatednear the ceiling surface of the combustion chamber 17 to be orientedtoward inside the combustion chamber 17. The electrode of the spark plug25 is adjacent to the nozzle ports of the injector 6. Note that thedisposed position of the spark plug 25 is not limited to theconfiguration example of FIG. 2. The spark plug 25 may be disposed onthe exhaust side of the axis X1. Alternatively, the spark plug 25 may bedisposed on the axis X1, and the injector 6 may be disposed on theintake side or the exhaust side with respect to the axis X1.

(Intake Passage 40)

An intake passage 40 is connected to one side of the engine 1. Theintake passage 40 communicates with the intake ports 18 of the cylinders11. The intake passage 40 is a passage through which gas to beintroduced into the combustion chamber 17 flows. An air cleaner 41 whichfilters fresh air is disposed in an upstream end part of the intakepassage 40. A surge tank 42 is disposed near a downstream end of theintake passage 40. Although not illustrated in detail, a part of theintake passage 40 downstream of the surge tank 42 constitutesindependent passages branched for the respective cylinders 11.Downstream ends of the independent passages communicate with the intakeports 18 of the cylinders 11, respectively.

A throttle valve 43 is disposed in the intake passage 40 between the aircleaner 41 and the surge tank 42. The throttle valve 43 changes anintroduction amount of fresh air into the combustion chamber 17 bychanging an opening thereof. The throttle valve 43 constitutes one ofthe state function setting devices.

A booster 44 is disposed in the intake passage 40 downstream of thethrottle valve 43. The booster 44 boosts the gas introduced into thecombustion chamber 17. In this configuration example, the booster 44 isa supercharger which is driven by the engine 1. The supercharger 44 maybe, for example, of a Lisholm type. The supercharger 44 may have anystructure. The supercharger 44 may be of a Roots type, a Vane type, or acentrifugal type. Note that the booster may be an electric booster or aturbocharger which is driven by the exhaust gas.

An electromagnetic clutch 45 is interposed between the booster 44 andthe engine 1. The electromagnetic clutch 45 controls the flow of adriving force between the booster 44 and the engine 1, for example, ittransmits the driving force from the engine 1 to the booster 44 orinterrupts the transmission of the driving force therebetween. As isdescribed later, by an ECU 10 switching the connection/disconnection ofthe electromagnetic clutch 45, the on/off of the booster 44 is switched.That is, in this engine 1, boosting the gas to be introduced into thecombustion chamber 17 by the booster 44 and not boosting the gas to beintroduced into the combustion chamber 17 by the booster 44 areswitchable therebetween.

An intercooler 46 is disposed in the intake passage 40 downstream of thebooster 44. The intercooler 46 cools the gas compressed in the booster44. The intercooler 46 may be, for example, of a water cooling type.

A bypass passage 47 is connected to the intake passage 40. The bypasspassage 47 connects a part of the intake passage 40 upstream of thebooster 44 to a part of the intake passage 40 downstream of theintercooler 46, particularly the surge tank 42. An air bypass valve 48is disposed in the bypass passage 47. The air bypass valve 48 changes aflow rate of the gas flowing through the bypass passage 47.

When the booster 44 is turned off (that is, when the electromagneticclutch 45 is disconnected), the air bypass valve 48 is fully opened.Thus, the gas flowing through the intake passage 40 bypasses the booster44 and is introduced into the combustion chamber 17 of the engine 1. Theengine 1 operates in a non-boosted state, that is, in a naturallyaspirated state.

When the booster 44 is turned on (that is, when the electromagneticclutch 45 is connected), it operates in a boosting state (a state wherepressure higher than atmospheric pressure is dynamically applied on thedownstream side of the booster 44). The gas passed through the booster44 partially flows back upstream of the booster 44 through the bypasspassage 47. By changing an opening of the air bypass valve 48, abackflow amount is changed, which leads to adjusting boosting pressureof the gas introduced into the combustion chamber 17. In thisconfiguration example, a boosting system 49 comprises the booster 44,the bypass passage 47, and the air bypass valve 48. The air bypass valve48 constitutes one of the state function setting devices.

(Swirl Control Valve 56)

Further in the intake passage 40, a swirl control valve (SCV) 56 (intakeflow control device) which controls a flow of intake air introduced intothe combustion chamber 17 to form a swirl flow therein and changes thestrength of the flow, is disposed.

As illustrated in FIG. 3, the SCV 56 is disposed in a secondary passage402. The secondary passage 402 is one of a primary passage 401 and thesecondary passage 402 communicating with the first intake port 181 andthe second intake port 182, respectively. The SCV 56 is an openingregulating valve which is capable of adjusting the opening of a crosssection of the secondary passage.

When the opening of the SCV 56 is small, the flow rate of the intake airinto the combustion chamber 17 from the first intake port 181 relativelyincreases while the flow rate of the intake air into the combustionchamber 17 from the second intake port 182 is relatively reduced. Thus,the swirl flow in the combustion chamber 17 becomes strong.

When the opening of the SCV 56 is large, the flow rates of the intakeair into the combustion chamber 17 from the first intake port 181 andthe second intake port 182 become substantially even, and thus the swirlflow in the combustion chamber 17 becomes weak. When the SCV 56 is fullyopened, a swirl flow does not occur. Note that the swirl flow circulatesin the counter-clockwise direction in FIG. 3 as indicated by the whiteoutlined arrows (also see the white outlined arrows in FIG. 2).

Note that alternatively/additionally to attaching the SCV 56 to theintake passage 40, a structure in which the open periods of the twointake valves 21 are varied so as to introduce the intake air into thecombustion chamber 17 from only one of the intake valves 21 may beadopted. By opening only one of the two intake valves 21, the intake airis introduced unevenly into the combustion chamber 17, and thus, theswirl flow is generated in the combustion chamber 17. Alternatively, theshapes of the intake ports 18 may be devised so that the swirl flow isgenerated in the combustion chamber 17.

(Swirl Flow)

Here, the strength of the swirl flow in the combustion chamber 17 isdefined. In this configuration example, the strength of the swirl flowin the combustion chamber 17 is expressed by a “swirl ratio.” The “swirlratio” may be defined as a value obtained by dividing a value which isobtained from measuring an intake flow lateral angular speed for eachvalve lift and integrating the value, by an engine angular speed. Theintake flow lateral angular speed may be obtained based on a measurementusing a rig test device illustrated in FIG. 5.

Specifically, in the rig test device illustrated in FIG. 5, the cylinderhead 13 is placed upside down on a pedestal to connect the intake port18 to an intake air supply device (not illustrated), and placing acylinder 36 on the cylinder head 13 to connect, at its upper end, to animpulse meter 38 having a honeycomb rotor 37. A lower surface of theimpulse meter 38 is positioned at a position 1.75D (wherein “D” is acylinder bore diameter) away from a mating surface between the cylinderhead 13 and the cylinder 36. The impulse meter 38 measures torque whichacts on the honeycomb rotor 37 by a swirl generated in the cylinder 36according to the supply of the intake air (see the arrow in FIG. 5), andthe intake flow lateral angular speed is obtained based on the torque.

FIG. 6 illustrates a relationship between the opening of the SCV 56 ofthe engine 1 and the swirl ratio. In FIG. 6, the opening of the SCV 56is expressed by an opening ratio with respect to the cross section ofthe secondary passage 402 when fully opened. The opening ratio of thesecondary passage 402 is 0% when the SCV 56 is fully closed, andincreases from 0% as the opening of the SCV 56 increases. The openingratio of the secondary passage 402 is 100% when the SCV 56 is fullyopened.

As illustrated in FIG. 6, in the engine 1, the swirl ratio becomesaround 6 when the SCV 56 is fully closed. In order to set the swirlratio to be 4 or higher, the opening of the SCV 56 may be adjustedwithin a range of the opening ratio of 0 to 15%.

Moreover, in order to set the swirl ratio to be less than 4, the openingof the SCV 56 may be adjusted within a range of the opening ratio below15%. Especially in order to secure the fluidity of the mixture gas andto control the stratification of the mixture gas in the combustionchamber 17 in cooperation with the SPCCI combustion within the operatingrange in which the engine load is low or medium, the swirl ratio ispreferably adjusted within a range of 1.5 to 3 (25% to 40% in terms ofthe opening of the SCV 56).

(Exhaust Passage 50)

An exhaust passage 50 is connected to a side of the engine 1 oppositefrom the intake passage 40. The exhaust passage 50 communicates with theexhaust ports 19 of the cylinders 11. The exhaust passage 50 is apassage through which the exhaust gas discharged from the combustionchamber 17 flows. Although not illustrated in detail, an upstream partof the exhaust passage 50 constitutes independent passages branched forthe respective cylinders 11. Upstream ends of the independent passagesare connected to the exhaust ports 19 of the cylinders 11, respectively.

An exhaust gas purification system having one or more catalyticconverters is disposed in the exhaust passage 50. The exhaust gaspurification system of this configuration example has two catalyticconverters. The catalytic converter on the upstream side is disposed inan engine bay and has a three-way catalyst 511 and a GPF (GasolineParticulate Filter) 512. The catalytic converter on the downstream sideis disposed outside the engine bay and has a three-way catalyst 513.

Note that the exhaust gas purification system is not limited to have theillustrated structure. For example, the GPF 512 may be omitted.Moreover, the catalytic converter is not limited to have the three-waycatalyst. Furthermore, the order of arrangements of the three-waycatalyst and the GPF may suitably be changed.

An EGR passage 52 constituting an external EGR system is connectedbetween the intake passage 40 and the exhaust passage 50. The EGRpassage 52 is a passage for recirculating a portion of the burned gas tothe intake passage 40. An upstream end of the EGR passage 52 isconnected to the exhaust passage 50 between the upstream catalyticconverter and the downstream catalytic converter. A downstream end ofthe EGR passage 52 is connected to the intake passage 40 upstream of thebooster 44.

For example, the downstream end of the EGR passage 52 is connected to anintermediate position of the bypass passage 47. The EGR gas flowingthrough the EGR passage 52 enters the intake passage 40 upstream of thebooster 44 without passing through the air bypass valve 48 of the bypasspassage 47.

A water-cooling type EGR cooler 53 is disposed in the EGR passage 52.The EGR cooler 53 cools the burned gas. An EGR valve 54 is also disposedin the EGR passage 52. The EGR valve 54 changes the flow rate of theburned gas in the EGR passage 52. By changing an opening of the EGRvalve 54, the recirculation amount of the cooled burned gas (i.e.,external EGR gas) is changed.

In this configuration example, an EGR system 55 includes an external EGRsystem including the EGR passage 52 and the EGR valve 54, and aninternal EGR system including the intake electrically-operated S-VT 23and the exhaust electrically-operated S-VT 24 described above. The EGRvalve 54 constitutes one of the state function setting devices. In theexternal EGR system, since the EGR passage 52 is connected downstream ofthe upstream catalytic converter and the EGR cooler 53 is provided, theburned gas at a temperature lower than in the internal EGR system issupplied to the combustion chamber 17.

(ECU 10)

A control system 20 of the engine 1 includes an ECU (Engine ControlUnit) 10 configured to operate the engine 1. The ECU 10 is a controllerbased on a well-known microcomputer. The ECU 10 includes a centralprocessing unit (CPU) 101 comprising a processor configured to executeprogram(s)/instructions, memory 102 comprised of RAM(s) (Random AccessMemory) and ROM(s) (Read Only Memory) and configured to store theprogram(s)/instructions and data, and an input/output bus 103 configuredto input and output electric signals.

The memory 102 stores a SI ratio changing module (heat amount ratiochanging module) 102 a which changes the SI ratio (heat amount ratio,described later in detail) according to the operating state of theengine 1 in order to control the SPCCI combustion. The SI ratio changingmodule 102 a comprises, for example, data such as a control program anda map used for the control program.

As illustrated in FIGS. 1 and 4, various sensors SW1 to SW16 areconnected to the ECU 10. The sensors SW1 to SW16 output detectionsignals to the ECU 10. The sensors include the following sensors.

That is, the sensors include an airflow sensor SW1 disposed in theintake passage 40 downstream of the air cleaner 41 and configured todetect the flow rate of fresh air in the intake passage 40, a firstintake air temperature sensor SW2 also disposed in the intake passage 40downstream of the air cleaner 41 and configured to detect a temperatureof the fresh air, a first pressure sensor SW3 disposed in the intakepassage 40 downstream of the connecting position with the EGR passage 52and upstream of the booster 44, and configured to detect pressure of thegas flowing into the booster 44, a second intake air temperature sensorSW4 disposed in the intake passage 40 downstream of the booster 44 andupstream of the connecting position of the bypass passage 47 andconfigured to detect a temperature of the gas flowed out of the booster44, a second pressure sensor SW5 attached to the surge tank 42 andconfigured to detect pressure of the gas at a position downstream of thebooster 44, pressure sensors SW6 attached to the cylinder head 13corresponding to the cylinders 11 and configured to detect pressure inthe combustion chambers 17, respectively, an exhaust temperature sensorSW7 disposed in the exhaust passage 50 and configured to detect atemperature of the exhaust gas discharged from the combustion chamber17, a linear O₂ sensor SW8 disposed in the exhaust passage 50 upstreamof the catalytic converter 511 and configured to detect an oxygenconcentration within the exhaust gas, a lambda O₂ sensor SW9 disposed inthe exhaust passage 50 downstream of the catalytic converter 511 andconfigured to detect an oxygen concentration within the exhaust gas, awater temperature sensor SW10 attached to the engine 1 and configured todetect a temperature of the cooling water, a crank angle sensor SW11attached to the engine 1 and configured to detect a rotational angle ofthe crankshaft 15, an accelerator opening sensor SW12 attached to anaccelerator pedal mechanism and configured to detect an acceleratoropening corresponding to an operation amount of an accelerator pedal, anintake cam angle sensor SW13 attached to the engine 1 and configured todetect a rotational angle of the intake camshaft, an exhaust cam anglesensor SW14 attached to the engine 1 and configured to detect arotational angle of the exhaust camshaft, an EGR pressure differencesensor SW15 disposed in the EGR passage 52 and configured to detect adifference in pressure between positions upstream and downstream of theEGR valve 54, and a fuel pressure sensor SW16 attached to the commonrail 64 of the fuel supply system 61 and configured to detect pressureof the fuel to be supplied to the injector 6.

Based on these detection signals, the ECU 10 determines the operatingstate of the engine 1 and calculates control amounts of the variousdevices. The ECU 10 outputs control signals related to the calculatedcontrol amounts to the injector 6, the spark plug 25, the intakeelectrically-operated S-VT 23, the exhaust electrically-operated S-VT24, the fuel supply system 61, the throttle valve 43, the EGR valve 54,the electromagnetic clutch 45 of the booster 44, the air bypass valve48, and the SCV 56.

For example, the ECU 10 changes the boosting pressure by changing theopening of the air bypass valve 48 based on a pressure differencebetween the upstream and downstream sides of the booster 44 obtainedfrom the detection signals of the first pressure sensor SW3 and thesecond pressure sensor SW5. Further, the ECU 10 changes the external EGRgas amount introduced into the combustion chamber 17 by changing theopening of the EGR valve 54 based on the pressure difference between theupstream and downstream sides of the EGR valve 54 obtained from thedetection signal of the EGR pressure difference sensor SW15. Details ofother controls of the engine 1 by the ECU 10 are described later.

<Operating Range of Engine>

FIG. 7A illustrates a first configuration example of an operating rangemap of the engine 1 (first operating range map 700). The operating rangemap 700 is defined by an engine load and an engine speed and is roughlydivided into the following four ranges based on the engine load and theengine speed.

(A): a low load range including an idle operation

(B): a medium load range between the low load range (A) and thefollowing high load range (C)

(C): a high load range including a full engine load

(D): a high-speed range where the engine speed is higher than in the lowload range (A), the medium load range (B), and the high load range (C)

The engine 1 performs the SPCCI combustion within the medium load range(B) in order to improve the fuel efficiency and exhaust gas performance.Hereinafter, the combustion modes in each of the low load range (A), themedium load range (B), and the high load range (C) will be described indetail.

(Low Load Range)

When the operating state of the engine 1 is within the low load range(A), the fuel injection amount is small. Therefore, the amount of heatgenerated when the mixture gas is combusted in the combustion chamber 17is small and the temperature of the combustion chamber 17 is low.Additionally, since the temperature of the exhaust gas is also low, evenif the internal EGR gas is introduced into the combustion chamber 17 asdescribed later, it is difficult to raise the temperature of thecombustion chamber 17 to such a degree that self-ignition is stablyperformed.

The combustion mode when the operating state of the engine 1 is withinthe low load range (A) is the SI combustion in which the spark plug 25ignites the mixture gas inside the combustion chamber 17 to combust itby flame propagation. Hereinafter, the combustion mode within the lowload range (A) may be referred to as “low-load SI combustion.”

When the operating state of the engine 1 is within the low load range(A), an air-fuel ratio (A/F) of the mixture gas is at the theoreticalair-fuel ratio (A/F≈14.7:1). Note that below, the air-fuel ratio, anexcess air ratio λ, and the value of G/F (gas/fuel ratio) of the mixturegas mean the values taken at an ignition timing.

When the air-fuel ratio of the mixture gas is set to the theoreticalair-fuel ratio, the three-way catalyst is able to purify the exhaust gasdischarged from the combustion chamber 17, and thus the exhaust gasperformance of the engine 1 improves. The A/F of the mixture gas may beset to remain within the purification window of the three-way catalyst(i.e., an air-fuel ratio width exhibiting the three-way purificationfunction). The excess air ratio λ of the mixture gas may be set to1.0±0.2.

In order to improve the fuel efficiency of the engine 1, when theoperating state of the engine 1 is within the low load range (A), theEGR system 55 introduces the EGR gas into the combustion chamber 17. TheG/F of the mixture gas, which is a mass ratio of the total gas to thefuel in the combustion chamber 17, is set between 18 and 30. The G/F ofthe mixture gas may be set between 18 and 50. The mixture gas is EGRlean and has a high dilution ratio.

By setting the G/F of the mixture gas to, for example, 25, within thelow load range (A), the SI combustion is stably performed without themixture gas self-igniting. Within the low load range (A), the G/F of themixture gas is maintained constant regardless of the engine load. Thus,the SI combustion is stable throughout the entire low load range.Additionally, the fuel efficiency of the engine 1 improves and theexhaust gas performance improves.

When the operating state of the engine 1 is within the low load range(A), since the fuel amount is low, a charge amount of gas into thecombustion chamber 17 needs to be lower than 100% in order to bring λ ofthe mixture gas to 1.0±0.2 and G/F to a value between 18 and 50. Forexample, the engine 1 executes throttling for changing the opening ofthe throttle valve 43 and/or a mirror cycle for retarding the closetiming of the intake valve 21 to after a bottom dead center (BDC) on theintake stroke.

In the engine 1 adopting the operating range map 700, when the operatingstate is within the low load range (A), the SCV 56 is substantiallyfully opened. Therefore, the swirl flow scarcely occurs in thecombustion chamber 17.

Note that within a low-load and low-speed segment of the low load range(A), the combustion temperature of the mixture gas and the temperatureof the exhaust gas may be raised by reducing the charge amount of gaseven smaller. This is advantageous in keeping the catalytic converter 51in an active state.

(Medium Load Range)

When the operating state of the engine 1 is within the medium load range(B), the fuel injection amount increases. The temperature of thecombustion chamber 17 increases, and thus, the self-ignition isperformed stably. Within the medium load range (B), the engine performsthe CI combustion in order to improve the fuel efficiency and exhaustgas performance.

In the combustion caused by self-ignition, the timing of theself-ignition changes greatly if the temperature inside the combustionchamber varies before the compression starts. Therefore, within themedium load range (B), the SPCCI combustion is performed.

Further, within the medium load range (B), the engine 1 sets the stateinside the combustion chamber 17 so that λ of the mixture gas becomes1.0±0.2 and the G/F of the mixture gas becomes a value between 18 and50. Moreover, at the ignition timing, a required temperature T_(IG)inside the combustion chamber 17 is 570 to 800K, a required pressureP_(IG) inside the combustion chamber 17 is 400 to 920 kPa, andturbulence kinetic energy inside the combustion chamber 17 is 17 to 40m²/s².

In the engine 1 adopting the operating range map 700, when the operatingstate is within the medium load range (B), the SCV 56 is substantiallyfully opened. Therefore, the swirl flow scarcely occurs in thecombustion chamber 17.

By accurately controlling the self-ignition timing, an increase of thecombustion noise is avoided when the operating state of the engine 1 iswithin the medium load range (B). Moreover, by increasing the dilutionratio of the mixture gas as high as possible and performing the CIcombustion, the fuel efficiency of the engine 1 is improved. Moreover,by setting λ of the mixture gas to 1.0±0.2, the three-way catalyst isable to purify the exhaust gas, and thus the exhaust gas performance ofthe engine 1 improves.

As described above, within the low load range (A), the G/F of themixture gas is set between 18 and 50 (e.g., 25) and λ of the mixture gasis set to 1.0±0.2. The state function inside the combustion chamber 17does not vary greatly between the states where the operating state ofthe engine 1 is within the low load range (A) and within the medium loadrange (B). Therefore, robustness of the control of the engine 1 againstthe change of the engine load improves.

When the operating state of the engine 1 is within the medium load range(B), different from being within the low load range (A), the fuel amountincreases, therefore the charge amount of gas introduced into thecombustion chamber 17 is not required to be changed. Here, the throttlevalve 43 is fully opened.

When the engine load increases and the fuel amount further increases, inthe naturally aspirated state, the introduction amount of gas into thecombustion chamber 17 becomes insufficient for setting λ of the mixturegas to 1.0±0.2 and the G/F of the mixture gas between 18 and 50.Therefore, in a segment of the medium load range (B) where the engineload is higher than a given load, the booster 44 boosts the gas to beintroduced into the combustion chamber 17.

The medium load range (B) is divided into a first medium load segment(B1) in which the engine load is higher than the given load and theboost is performed, and a second medium load segment (B2) in which theengine load is lower than the given load and the boost is not performed.The given load is, for example, ½ load. The second medium load segment(B2) is a segment where the engine load is lower than the first mediumload segment (B1). Hereinafter, the combustion mode within the firstmedium load segment (B1) may be referred to as “boosted SPCCIcombustion” and the combustion mode within the second medium loadsegment (B2) may be referred to as “non-boosted SPCCI combustion.”

Within the second medium load segment (B2) in which the boost is notperformed, as the fuel amount increases, the introduction amount offresh air into the combustion chamber 17 increases while the EGR gasdecreases. The G/F of the mixture gas decreases as the engine loadincreases. Since the throttle valve 43 is fully opened, the engine 1changes the introduction amount of EGR gas into the combustion chamber17 to change the amount of fresh air introduced into the combustionchamber 17. Within the second medium load segment (B2), the statefunction inside the combustion chamber 17 is set such that, for example,λ of the mixture gas is substantially constant at 1.0 while the G/F ofthe mixture gas is changed between 25 and 28.

On the other hand, within the first medium load segment (B1) in whichthe boost is performed, the engine 1 increases the introduction amountsof fresh air and EGR gas into the combustion chamber 17 as the fuelamount increases. Thus, the G/F of the mixture gas is constant even whenthe engine load increases. In the state function inside the combustionchamber 17 within the first medium load segment (B1), for example, λ ofthe mixture gas is substantially constant at 1.0 and the G/F of themixture gas is constant at 25.

(High Load Range)

The combustion mode when the operating state of the engine 1 is withinthe high load range is the SI combustion. This is for prioritizingreliably avoiding the combustion noise. Hereinafter, the combustion modewithin the high load range may be referred to as “high-load SIcombustion.”

When the operating state of the engine is within the high load range(C), λ of the mixture gas becomes 1.0±0.2, and the G/F of the mixturegas is basically set at between 18 and 30. The G/F of the mixture gasmay be set between 18 and 50. Within the high load range (C), thethrottle valve 43 is fully opened and the booster 44 performs the boost.

Within the high load range (C), the engine 1 reduces the EGR gas amountas the engine load increases. The G/F of the mixture gas decreases asthe engine load increases. The introduction amount of fresh air into thecombustion chamber 17 increases by the reduced amount of EGR gas,therefore, the fuel amount may be increased, which is advantageous inincreasing a highest output of the engine 1.

In the engine 1 adopting the operating range map 700, when the operatingstate is within the high load range (C), the SCV 56 is substantiallyfully opened. Therefore, the swirl flow scarcely occurs in thecombustion chamber 17.

The state function inside the combustion chamber 17 does not varygreatly between the states where the operating state of the engine 1 iswithin the high load range (C) and within the medium load range (B).Therefore, the robustness of the control of the engine 1 against thechange of the engine load improves.

Since the engine 1 performs the SI combustion within the high load range(C) as described above, there is an issue with abnormal combustion, suchas pre-ignition and knocking, occurring easily.

Therefore, within the high load range (C), by devising the fuelinjection mode, abnormal combustion is avoided in the engine 1. Forexample, the ECU 10 outputs control signals to the fuel supply system 61and the injector 6 to inject the fuel into the combustion chamber 17 ata high fuel pressure of 30 MPa or higher, at a timing in a period from afinal stage of the compression stroke to an early stage of the expansionstroke (hereinafter, this period is referred to as “retard period”). TheECU 10 also outputs a control signal to the spark plug 25 to ignite themixture gas at a timing near CTDC after the fuel injection. Hereinafter,the fuel injection into the combustion chamber 17 at the high fuelpressure at the timing in the retard period is referred to as“high-pressure retard injection.”

The high-pressure retard injection shortens reaction time of the mixturegas to avoid abnormal combustion. That is, the reaction time of themixture gas is a total length of time of (1) a period for which theinjector 6 injects the fuel (i.e., injection period), (2) a period forwhich combustible mixture gas is formed around the spark plug 25 afterthe fuel injection (i.e., mixture gas formation period), and (3) aperiod from the start of ignition until the SI combustion ends (i.e.,(3) combustion period).

The injection period and the mixture gas formation period become shorterby injecting the fuel into the combustion chamber 17 at the high fuelpressure. By shortening the injection period and the mixture gasformation period, the timing of starting the fuel injection approachesthe ignition timing. In the high-pressure retard injection, since thefuel injection into the combustion chamber 17 is performed at the highfuel pressure, the fuel is injected at a timing in the retard periodfrom the final stage of the compression stroke to the early stage of theexpansion stroke.

Injecting the fuel into the combustion chamber 17 at the high fuelpressure increases turbulence kinetic energy inside the combustionchamber 17. By bringing the fuel injection timing close to CTDC, it ispossible to start the SI combustion while the turbulence kinetic energyinside the combustion chamber 17 is high. As a result, the combustionperiod becomes short.

Thus, in the high-pressure retard injection, since the injection period,the mixture gas formation period, and the combustion period arerespectively shortened, the reaction time of the mixture gas issignificantly shortened compared with a case where the fuel is injectedinto the combustion chamber 17 on the intake stroke. As a result,abnormal combustion is avoided.

In the technical field of the engine control, conventionally, theignition timing is retarded to avoid abnormal combustion. However,retarding the ignition timing degrades the fuel efficiency. In thehigh-pressure retard injection, the ignition timing is not required tobe retarded. Therefore, the fuel efficiency improves by using thehigh-pressure retard injection.

By setting the fuel pressure to be, for example, 30 MPa or higher, theinjection period, the mixture gas formation period, and the combustionperiod are effectively shortened. Note that the fuel pressure maysuitably be set according to properties of the fuel. An upper limit ofthe fuel pressure may be, for example, 120 MPa.

Here, when the engine speed is low, compared to when it is high, thetime required for the crank angle to change by the same angle is longer,therefore, shortening the reaction time of the mixture gas by thehigh-pressure retard injection is particularly effective in avoidingabnormal combustion. On the other hand, when the engine speed is high,due to the shorter time required for the crank angle to change by thesame angle, shortening the reaction time of the mixture gas is notparticularly effective in avoiding abnormal combustion.

Further in the high-pressure retard injection, the fuel is injected intothe combustion chamber 17 only after reaching near CTDC, on thecompression stroke, fuel-free gas, in other words, the gas with a highratio of specific heat, is compressed within the combustion chamber 17.If the high-pressure retard injection is performed when the engine speedis high, the temperature inside the combustion chamber 17 at CTDC, i.e.,the compression end temperature, rises, which may cause abnormalcombustion, such as knocking.

Therefore, in the engine 1, the high load range (C) is divided into afirst high load segment (C1) on the low engine speed side and a secondhigh load segment (C2) where the engine speed is higher than within thefirst high load segment (C1). When the high load range (C) is evenlydivided into three ranges of low engine speed, medium engine speed, andhigh engine speed, the first high load segment (C1) may include the lowengine speed range and the medium engine speed range, and the secondhigh load segment (C2) may include the high engine speed range.

Within the first high load segment (C1), the injector 6, in response toreceiving the control signal of the ECU 10, performs the high-pressureretard injection described above. Within the second high load segment(C2), the injector 6, in response to receiving the control signal of theECU 10, performs the fuel injection at a given timing on the intakestroke. The fuel injection performed on the intake stroke does notrequire high fuel pressure. Therefore, the ECU 10 outputs the controlsignal to the fuel supply system 61 so that the fuel pressure fallsbelow the fuel pressure of the high-pressure retard injection (e.g.,below 40 MPa). Since lowering the fuel pressure reduces a mechanicalresistance loss of the engine 1, it is advantageous in improving thefuel efficiency.

The ratio of the specific heat of the gas inside the combustion chamber17 decreases by injecting the fuel into the combustion chamber 17 on theintake stroke, therefore, the compression end temperature drops, andthus, abnormal combustion is avoided. Since it is not necessary toretard the ignition timing for avoiding abnormal combustion, within thesecond high load segment (C2), similar to the first high load segment(C1), the spark plug 25 ignites the mixture gas at a timing near CTDC.

Within the first high load segment (C1), since the mixture gas does notresult in self-ignition because the high-pressure retard injection isapplied, the engine 1 performs stable SI combustion. Within the secondhigh load segment (C2), since the mixture gas does not result inself-ignition because the fuel is injected on the intake stroke, theengine 1 performs stable SI combustion.

<SPCCI Combustion>

Next, the SPCCI combustion described above is described in detail. Theupper chart of FIG. 8 illustrates a waveform 801 which is one example ofa change in a heat generation rate with respect to the crank angle. Whenthe spark plug 25 ignites the mixture gas near CTDC, specifically at agiven timing before CTDC, the combustion starts by flame propagation.The heat generation in the SI combustion is slower than the heatgeneration in the CI combustion. Therefore, the waveform of the heatgeneration rate has a relatively gentle slope. Although not illustrated,a pressure variation (dp/dθ) in the combustion chamber 17 is gentler inthe SI combustion than in the CI combustion.

When the temperature and pressure inside the combustion chamber 17 risedue to the SI combustion, the unburned mixture gas self-ignites. In theexample of the waveform 801, the slope of the waveform of the heatgeneration rate changes from gentle to sharp, i.e., the waveform of theheat generation rate has a flexion point at a timing when the CIcombustion starts.

After the CI combustion starts, the SI combustion and the CI combustionare performed in parallel. In the CI combustion, since the heatgeneration is greater than in the SI combustion, the heat generationrate becomes relatively high. Note that since the CI combustion isperformed after CTDC, the piston 3 descends by motoring. Therefore, theslope of the waveform of the heat generation rate by the CI combustionis avoided from becoming excessively sharp. The dp/dθ in the CIcombustion also becomes relatively gentle.

The dp/dθ is usable as an index expressing the combustion noise. Sincethe SPCCI combustion is able to lower the dp/dθ as described above, itbecomes possible to avoid the combustion noise from becoming excessivelyloud. Thus, the combustion noise is suppressed below an allowable value.

The SPCCI combustion ends by finishing the CI combustion. The CIcombustion has a shorter combustion period than in the SI combustion.The SPCCI combustion advances the combustion end timing compared to theSI combustion. In other words, the SPCCI combustion brings thecombustion end timing on the expansion stroke closer to CTDC. The SPCCIcombustion is advantageous in improving the fuel efficiency of theengine 1 than the SI combustion.

Therefore, the SPCCI combustion achieves both prevention of thecombustion noise and improvement in the fuel efficiency.

(SI Ratio:Heat Amount Ratio)

Here, an SI ratio is defined as a parameter indicating a characteristicof the SPCCI combustion. The SI ratio is defined as an index relating toa ratio of the heat amount generated by the SI combustion with respectto a total heat amount generated by the SPCCI combustion. The SI ratiois a heat amount ratio resulted from two combustions with differentcombustion modes. The SI ratio may be a ratio of heat amount generatedby the SI combustion with respect to the heat amount generated by theSPCCI combustion. For example, in the waveform 801, the SI ratio may beexpressed as SI ratio=(area of SI combustion)/(area of SPCCIcombustion). In the waveform 801, the SI ratio may be referred to as “SIfuel ratio” in the meaning of the ratio of fuel to be combusted in theSI combustion.

In the SPCCI combustion combined the SI combustion and the CIcombustion, the SI ratio is a ratio of the SI combustion with respect tothe CI combustion. The ratio of the SI combustion is high when the SIratio is high, and the ratio of the CI combustion is high when the SIratio is low.

Various definitions may be considered for the SI ratio without limitingto the definition described above. For example, the SI ratio may be aratio of the heat amount generated by the SI combustion with respect tothe heat amount generated by the CI combustion. In other words, in FIG.8, SI ratio=(area of SI combustion)/(area of CI combustion) may be set.

Further, in the SPCCI combustion, the waveform of the heat generationrate has a flexion point at the timing when the CI combustion starts.Therefore, as indicated by a reference character 802 in the middle chartof FIG. 8, by having a boundary at the flexion point in the waveform ofthe heat generation rate, the SI combustion may be applied for a rangeon the advancing side of the boundary, and the CI combustion may beapplied for a range on the retarding side of the boundary. In this case,as indicated by hatching the waveform 802, based on an area Q_(SI) ofthe advancing-side range and an area Q_(CI) of the retarding-side range,SI ratio=Q_(SI)/(Q_(SI)+Q_(CI)) or SI ratio=Q_(SI)/Q_(CI) may be set.Alternatively, the SI ratio may be defined based on an area of a part ofthe advancing-side range and an area of a part of the retarding-siderange, instead of the entire area.

Further, instead of defining the SI ratio based on the heat generation,based on a crank angle Δθ_(SI) of the advancing-side range and a crankangle Δθ_(CI) of the retarding-side range, SIratio=Δθ_(SI)/(Δθ_(SI)+Δθ_(CI)) or SI ratio=Δθ_(SI)/Δθ_(CI) may be set.

Moreover, based on a peak ΔP_(SI) of the heat generation rate in theadvancing-side range and a peak ΔP_(CI) of the heat generation rate inthe retarding-side range, SI ratio=ΔP_(SI)/(ΔP_(SI)+ΔP_(CI)) or SIratio=ΔP_(SI)/ΔP_(CI) may be set.

Furthermore, based on a slope φ_(SI) of the heat generation rate in theadvancing-side range and a slope φ_(CI) of the heat generation rate inthe retarding-side range, SI ratio=φ_(SI)/(φ_(SI)+φ_(CI)) or SIratio=φ_(SI)/φ_(CI) may be set.

Additionally, in this embodiment, the SI ratio is defined by one of thearea (i.e., the heat generation amount), length in the horizontal axis(i.e., the crank angle), length in the vertical axis (i.e., the heatgeneration rate), and the slope (i.e., the change rate in the heatgeneration rate) based on the waveform of the heat generation rate.Although not illustrated, the SI ratio may similarly be defined based ona waveform of pressure (P) in the combustion chamber 17, by one of thearea, length in the horizontal axis, length in the vertical axis, andthe slope.

In the SPCCI combustion, the flexion point of the combustion waveformregarding the heat generation rate or pressure does not necessarilyappear clearly all the time. The following may be used as a definitionof the SI ratio which is not based on the flexion point. That is, asindicated by a reference character 803 in the lower chart of FIG. 8, inthe combustion waveform, the SI combustion may be applied for a range onthe advancing side of CTDC and the CI combustion may be applied for arange on the retarding side of CTDC. Under this condition, the SI ratiomay be defined by one of the area (Q_(SI), Q_(CI)), length in thehorizontal axis (Δθ_(SI), Δθ_(CI)), length in the vertical axis(ΔP_(SI), ΔP_(CI)) and the slope (φ_(SI), φ_(CI)).

Alternatively, the SI ratio may be defined based on the fuel amountinstead of the actual combustion waveform in the combustion chamber 17.As described later, within the medium load range in which the SPCCIcombustion is performed, split injections including a first-stageinjection and a second-stage injection may be performed. The fuelinjected into the combustion chamber 17 by the second-stage injectionignites within a short time after the injection, it reaches near thespark plug 25 without spreading inside the combustion chamber 17.Therefore, the fuel injected into the combustion chamber 17 by thesecond-stage injection combusts mainly in the SI combustion.

On the other hand, the fuel injected into the combustion chamber 17 bythe first-stage injection combusts mainly in the CI combustion.Therefore, the SI ratio may be defined based on the fuel amount injectedin the first-stage injection (m₁) and the fuel amount injected in thesecond-stage injection (m₂). In other words, SI ratio=m₂/(m₁+m₂) or SIratio=m₂/m₁ may be set.

(Stabilizing SPCCI Combustion)

In order to appropriately perform the SPCCI combustion, the SIcombustion needs to be stabilized. The combustion which includes the CIcombustion does not stabilize if the SI combustion is unstable.

One factor related to the stability of the SI combustion is a turbulentcombustion rate. When the turbulent combustion rate is high, the SIcombustion stabilizes. The turbulent combustion rate receives influencesfrom the air-fuel ratio (or the excess air ratio k) of the mixture gas,the G/F of the mixture gas, the temperature and pressure in thecombustion chamber, the turbulence kinetic energy in the combustionchamber, etc.

According to the study of the present inventors, it was confirmed thatthe SI combustion stabilizes when λ of the mixture gas is 1.0±0.2, therange of G/F of the mixture gas is 18 to 30, the range of the requiredtemperature T_(IG) inside the combustion chamber at the ignition timingis 570 to 800K, the range of the required pressure P_(IG) inside thecombustion chamber at the ignition timing is 400 to 920 kPa, and therange of the turbulence kinetic energy in the combustion chamber is 17to 40 m²/m².

It was also confirmed that it is possible to extend the range of G/F ofthe mixture gas to 50 which is above 30, by stratifying the mixture gasinside the combustion chamber.

That is, the SI combustion in the SPCCI combustion is caused by thespark plug 25 igniting the mixture gas. The mixture gas near the sparkplug 25 mainly combusts in the SI combustion. Therefore, the state ofthe fuel spray moving in the combustion chamber 17 is controlled by, forexample, utilizing the swirl flow. In this manner, even if the range ofG/F in the combustion chamber exceeds 30 at the ignition timing, the G/Fof the mixture gas near the spark plug 25 is made relatively lower(stratified) than the mixture gas located away from the spark plug 25,such as between 14 and 22.

Thus, the SI combustion is stabilized even with the mixture gas dilutedadvantageously for improving the fuel efficiency, and the SPCCIcombustion is appropriately performed.

<Operation Control of Engine in Load Direction>

By adopting the operating range map 700, the engine 1 switches betweenthe SI combustion and the SPCCI combustion according to the operatingstate. Further, the engine 1 changes the SI ratio according to theoperating state of the engine 1. Thus, the engine 1 is achieved inpreventing the generation of combustion noise and improving the fuelefficiency.

FIG. 9 is a chart illustrating a change in the SI ratio, a change in thestate function inside the combustion chamber 17, changes in the openperiods of the intake valve 21 and the exhaust valve 22, and changes inthe injection timing and ignition timing of the fuel, with respect tothe engine load. Hereinafter, the operation control of the engine 1 isdescribed for a condition in which the engine load gradually increasesat a given engine speed.

(Low Load Range (Low-Load SI Combustion))

Within the low load range (A), the engine 1 performs the low-load SIcombustion. When the operating state of the engine 1 is within the lowload range (A), the SI ratio is constant at 100%.

Within the low load range (A), as described above, the G/F of themixture gas is fixed between 18 and 50. The engine 1 introduces thefresh air and the burned gas by amounts corresponding to the fuelamount, into the combustion chamber 17. The introduction amount of freshair, as described above, is changed by throttling and/or the mirrorcycle. Since the dilution ratio is high, the temperature inside thecombustion chamber 17 is raised to stabilize the SI combustion. Withinthe low load range (A), the engine 1 introduces the internal EGR gasinto the combustion chamber 17.

The internal EGR gas is introduced into the combustion chamber 17 (i.e.,the burned gas is confined inside the combustion chamber 17) byproviding a negative overlap period in which the intake and exhaustvalves 21 and 22 are both closed over the exhaust TDC. The change of theinternal EGR gas amount is performed by suitably setting the length ofthe negative overlap period by the intake electrically-operated S-VT 23changing the open timing of the intake valve 21 and the exhaustelectrically-operated S-VT 24 changing the close timing of the exhaustvalve 22.

Note that the internal EGR gas may be introduced into the combustionchamber 17 by providing a positive overlap period in which the intakeand exhaust valves 21 and 22 are both opened.

That is, a period during which the open period of the intake valve 21and the open period of the exhaust valve 22 overlap is provided. In thismanner, high-temperature burned gas in the combustion chamber 17partially flows into the intake port 18 communicating with thecombustion chamber 17, by opening the intake valve 21. The burned gasflowed into the intake port 18 is introduced again into the combustionchamber 17 on the intake stroke. The amount of internal EGR gas ischangeable by changing the length of the positive overlap period and thetiming of the positive overlap period.

Note that the introduction of the internal EGR gas by setting thepositive overlap period advances the open period of the intake valve 21compared to the introduction of the internal EGR gas by setting thenegative overlap period. Advancing the open period of the intake valve21 increases the effective compression ratio and raise the temperaturein the combustion chamber 17. Therefore, it becomes possible to set alow geometric compression ratio of the engine 1 while securing thestability of the SPCCI combustion corresponding to the temperature rise.Thus, the cooling loss, the mechanical loss, a pumping loss, etc. areachieved.

Further, the introduction of the internal EGR gas by setting thepositive overlap period lowers the temperature of the combustion chambermore than when the internal EGR gas is introduced by setting thenegative overlap period in which the high-temperature burned gas isconfined as it is in the combustion chamber. As a result, theself-ignition in the SPCCI combustion is subsided and excessive CIcombustion is prevented.

Within the low load range (A), the charge amount into the combustionchamber 17 is changed to be below 100%. The amount of fresh airintroduced into the combustion chamber 17 and the amount of the internalEGR gas gradually increase as the fuel amount increases. The EGR ratiowithin the low load range (A) is, for example, 40%.

The injector 6 injects the fuel into the combustion chamber 17 on theintake stroke. Inside the combustion chamber 17, homogeneous mixture gasin which the excess air ratio λ is 1.0±0.2 and the G/F is between 18 and50 is formed. The excess air ratio λ is preferably 1.0 to 1.2. By thespark plug 25 igniting the mixture gas at the given timing before CTDC,the mixture gas combusts by flame propagation without reaching theself-ignition.

(Second Medium Load Segment (Non-Boosted SPCCI Combustion))

When the engine load increases and the operating state enters the secondmedium load segment (B2), the engine 1 switches from the low-load SIcombustion to the non-boosted SPCCI combustion. The SI ratio falls below100%. The fuel amount increases as the engine load increases. When theengine load is low within the second medium load segment (B2), the ratioof the CI combustion is increased as the fuel amount increases. The SIratio gradually decreases as the engine load increases. In the exampleof FIG. 9, the SI ratio decreases to a given load (lowest value) G1lower than 50%.

Since the fuel amount increases, the combustion temperature rises withinthe second medium load segment (B2). If the temperature inside thecombustion chamber 17 rises excessively, the heat generation at thestart of the CI combustion becomes sharp, which results in increasingthe combustion noise.

Therefore, within the second medium load segment (B2), the ratio betweenthe internal EGR gas and the external EGR gas is changed according tothe change in the engine load in order to change the temperature insidethe combustion chamber 17 before the compression starts. That is, as theengine load increases, the internal EGR gas is gradually reduced and thecooled external EGR gas is gradually increased. Within the second mediumload segment (B2), the negative overlap period changes from a longestlength to zero as the engine load increases. Also within the secondmedium load segment (B2), the internal EGR gas becomes zero when theengine load reaches a highest value.

This is similar for when the positive overlap period between the intakeand exhaust valves 21 and 22 is provided to introduce the internal EGRgas into the combustion chamber 17. As a result of adjusting thetemperature inside the combustion chamber 17 by adjusting the overlapperiod, the SI ratio in the SPCCI combustion is adjusted.

Within the second medium load segment (B2), the opening of the EGR valve54 is changed so that the external EGR gas increases as the engine loadincreases. The amount of external EGR gas introduced into the combustionchamber 17 is changed, when expressed by the EGR ratio, between 0 and30%, for example. Within the second medium load segment (B2), the EGRgas is replaced from the internal EGR gas to the external EGR gas as theengine load increases. Since the temperature inside the combustionchamber 17 is also controlled by adjusting the EGR ratio, the SI ratioof the SPCCI combustion is changed.

Note that the EGR gas amount introduced into the combustion chamber 17is continuous between the low load range (A) and the second medium loadsegment (B2). Within a low engine load section of the second medium loadsegment (B2), a large amount of the internal EGR gas is introduced intothe combustion chamber 17 same as the low load range (A). Since thetemperature inside the combustion chamber 17 becomes high, the mixturegas surely self-ignites when the engine load is low. Within a highengine load section of the second medium load segment (B2), the externalEGR gas is introduced into the combustion chamber 17. Since thetemperature inside the combustion chamber 17 decreases, the combustionnoise accompanying the CI combustion is reduced when the engine load ishigh.

Within the second medium load segment (B2), the charge amount introducedinto the combustion chamber 17 is set to 100%. The throttle valve 43 isfully opened. By changing the EGR gas amount which is a total of theinternal EGR gas and the external EGR gas, the introduction amount offresh air into the combustion chamber 17 is changed to the amountcorresponding to the fuel amount.

As the ratio of the CI combustion in the non-boosted SPCCI combustionincreases, the self-ignition timing is advanced. If the self-ignitiontiming is advanced than CTDC, the heat generation at the start of the CIcombustion becomes sharp, which results in increasing the combustionnoise. Therefore, in the engine 1, once the engine load reaches thegiven load G1, the SI ratio is gradually increased as the engine loadfurther increases therefrom.

That is, the engine 1 increases the ratio of the SI combustion as thefuel amount increases. For example, as illustrated in the upper chart ofFIG. 10, in the non-boosted SPCCI combustion, the ignition timing isgradually advanced as the fuel amount increases. Since the temperatureinside the combustion chamber 17 is changed by reducing the introductionamount of the internal EGR gas and increasing the introduction amount ofthe external EGR gas as described above, the temperature rise at CTDC issuppressed even when the SI ratio is increased as the fuel amountincreases. The slope of the heat generation rate of the SI combustionscarcely changes even when the engine load increases. Advancing theignition timing causes the SI combustion to start earlier, and the heatgeneration amount of SI combustion accordingly increases.

As a result of suppressing the temperature rise inside the combustionchamber 17 caused by the SI combustion, the unburned mixture gasself-ignites at a timing after CTDC. The heat generation amount by theCI combustion is substantially the same even when the engine loadincreases, since the heat generation amount of the SI combustion isincreased. Therefore, by setting the SI ratio to be gradually higheraccording to the increase in the engine load, the combustion noise isavoided from increasing. Note that the center of gravity of combustionin the non-boosted SPCCI combustion retards as the engine loadincreases.

Within the second medium load segment (B2), on the compression stroke,the injector 6 injects the fuel into the combustion chamber 17 in twostages, the first-stage injection and the second-stage injection. In thefirst-stage injection, the fuel is injected at a timing separated fromthe ignition timing, and in the second-stage injection, the fuel isinjected at a timing close to the ignition timing.

(First Medium Load Segment (Boosted SPCCI Combustion))

When the engine load further increases and the operating state of theengine 1 enters the first medium load segment (B1), the booster 44boosts the fresh air and the external EGR gas. The amount of fresh airintroduced into the combustion chamber 17 and the amount of the externalEGR gas both increase as the engine load increases. The amount ofexternal EGR gas introduced into the combustion chamber 17 is, whenexpressed by the EGR ratio, 30%, for example. The EGR ratio issubstantially constant regardless of the engine load. Therefore, the G/Fof the mixture gas is also substantially constant regardless of theengine load. Note that the EGR gas amount introduced into the combustionchamber 17 is continuous between the second medium load segment (B2) andthe first medium load segment (B1).

The SI ratio is fixed at a given value below 100% with respect to theengine load. When the SI ratio of the second medium load segment (B2),particularly the SI ratio which gradually increases as the engine loadincreases from a value above the given load G1 is compared with the SIratio of the first medium load segment (B1), the SI ratio of the firstmedium load segment (B1) where the engine load is higher is higher thanthat of the second medium load segment (B2). The SI ratio is continuousover the boundary between the first medium load segment (B1) and thesecond medium load segment (B2).

Here, within the first medium load segment (B1), the SI ratio mayslightly be changed according to the change of the engine load. Thechange rate of the SI ratio according to the engine load within thefirst medium load segment (B1) may be lower than that at a high engineload side of the second medium load segment (B2).

As illustrated in the lower chart of FIG. 10, also in the boosted SPCCIcombustion, the ignition timing is gradually advanced as the fuel amountincreases. Since the fresh air and the EGR gas amount introduced intothe combustion chamber 17 are increased by boosting as described above,the heat amount is large. Therefore, even when the fuel amountincreases, the temperature increase inside the combustion chamber 17caused by the SI combustion is suppressed. The waveform of the heatgeneration rate of the boosted SPCCI combustion becomes larger (the areaof the section formed by the waveform and the horizontal axis becomeslarger) in a similar shape as the engine load increases.

That is, the heat generation amount of the SI combustion increases whilethe slope of the heat generation rate of the SI combustion scarcelychanges. The unburned mixture gas self-ignites at substantially the sametiming after CTDC. The heat generation amount of the CI combustionincreases as the engine load increases. As a result, within the firstmedium load segment (B1), since both the heat generation amount of theSI combustion and the heat generation amount of the CI combustionincrease, the SI ratio is constant with respect to the engine load.Although the combustion noise increases when the peak of the heatgeneration of the CI combustion rises, since the engine load isrelatively high within the first medium load segment (B1), a certainlevel of combustion noise is allowed. Note that the center of gravity ofcombustion in the boosted SPCCI combustion retards as the engine loadincreases.

Within the first medium load segment (B1), the overlap period in whichthe intake and exhaust valves 21 and 22 both open (positive overlapperiod) is provided over the exhaust TDC. The unburned gas residing inthe combustion chamber 17 is scavenged by the boosting pressure. Thus,the temperature in the combustion chamber 17 drops, and as a result,abnormal combustion is prevented from occurring when the engine load isrelatively high. Further, by dropping the temperature in the combustionchamber 17, within the range where the engine load is relatively high,the self-ignition timing is adjusted to a suitable timing and the SIratio is maintained at a given SI ratio. Further, by scavenging theburned gas, the charge amount of fresh air in the combustion chamber 17is increased.

Within the first medium load segment (B1), similarly to the secondmedium load segment (B2), the injector 6 injects the fuel into thecombustion chamber 17 in two stages, the first-stage injection and thesecond-stage injection on the compression stroke. In the first-stageinjection, the fuel is injected at the timing separated from theignition timing, and in the second-stage injection, the fuel is injectedat the timing close to the ignition timing. The first-stage injectionmay be performed in the period from the intake stroke to the early halfof the compression stroke, and the second-stage injection may beperformed in the period from the latter half of the compression stroketo the early half of the expansion stroke.

By the injector 6 performing the first-stage injection and thesecond-stage injection in these periods, in the combustion chamber 17,substantially homogeneous mixture gas in which the excess air ratio λ is1.0±0.2 and the G/F is 18 to 30 is formed as a whole. Since the mixturegas is substantially homogeneous, the improvement in the fuel efficiencyby reducing an unburned fuel loss and the improvement in the exhaust gasperformance by avoiding the smoke generation are achieved. The excessair ratio λ is preferably 1.0 to 1.2.

By the spark plug 25 igniting the mixture gas at the given timing beforeCTDC, the mixture gas combusts by flame propagation. After thiscombustion starts, the unburned mixture gas self-ignites and causes theCI combustion. The fuel injected in the second-stage injection mainlycauses the SI combustion. The fuel injected in the first-stage injectionmainly causes the CI combustion. Since the first-stage injection isperformed on the compression stroke, the fuel injected in thefirst-stage injection is prevented from causing abnormal combustion,such as the pre-ignition. Moreover, the fuel injected in thesecond-stage injection is stably combusted by flame propagation.

(High Load Range (High-Load SI Combustion))

When the engine load further increases and the operating state of theengine 1 enters the high load range (C), the engine 1 performs thehigh-load SI combustion. Therefore, the SI ratio within the high loadrange (C) becomes 100%.

The throttle valve 43 is fully opened. The booster 44 boosts the freshair and the external EGR gas even within the high load range (C). TheEGR valve 54, by changing its opening, gradually reduces theintroduction amount of the external EGR gas as the engine loadincreases. In this manner, the fresh air introduced into the combustionchamber 17 increases as the engine load increases. Since the fuel amountis increased as the fresh air amount increases, it is advantageous inincreasing the highest output of the engine 1. Note that the EGR gasamount introduced into the combustion chamber 17 is continuous betweenthe first medium load segment (B1) and the high load range (C).

Also within the high load range (C), similarly to the first medium loadsegment (B1), the overlap period in which the intake and exhaust valves21 and 22 are both opened is provided over the exhaust TDC. The unburnedgas residing in the combustion chamber 17 is scavenged by the boostingpressure. Thus, the occurrence of abnormal combustion is prevented.Further, the charge amount of fresh air in the combustion chamber 17 isincreased.

Within a low engine speed segment of the high load range (C) (i.e., thefirst high load segment (C1)), the injector 6 injects the fuel into thecombustion chamber 17 in the retard period as described above. Within ahigh engine speed segment of the high load range (C) (i.e., the secondhigh load segment (C2)), the injector 6 injects the fuel into thecombustion chamber 17 on the intake stroke. Within either segment,substantially homogeneous mixture gas in which the excess air ratio λ is1.0±0.2 and the G/F is 18 to 50 is formed inside the combustion chamber17.

At the highest load, the excess air ratio λ is 0.8, for example.Moreover, the G/F of the mixture gas may be, for example, 17 at thehighest load. By the spark plug 25 igniting the mixture gas at the giventiming before CTDC, the mixture gas is combusted by flame propagation.Within the high load range (C), due to the high-pressure retardinjection or the fuel injection on the intake stroke, the mixture gascauses the SI combustion without reaching the self-ignition.

<Operation Control of Engine in Speed Direction>

(Adjusting SI Ratio)

FIG. 11 illustrates a relationship between the engine speed and the SIratio at a given load within the medium load range (B) in which theSPCCI combustion is performed. As illustrated in the operating range map700 of FIG. 7A, a speed N1 is a lowest speed within the medium loadrange (B). A speed N2 is a highest speed within the medium load range(B), located at a boundary with the high speed range (D) in which the SIcombustion is performed (same in the description below).

The CI combustion is disadvantageous in that the combustion noise islouder than in the SI combustion. However, when the engine speed is low,NVH of the engine 1 is generally low. Therefore, at the low enginespeed, even with the addition of the combustion noise of the CIcombustion to NVH, NVH is kept lower than an allowable value. Therefore,at the low engine speed, the ECU 10 lowers the SI ratio and the CIcombustion is sufficiently performed. Thus, at the low engine speed, thefuel efficiency is improved without NVH becoming a problem.

On the other hand, when the engine speed increases, NVH of the engine 1also increases accordingly. Moreover, with the addition of thecombustion noise of the CI combustion on top of the increased NVH, NVHmay exceed the allowable value. Therefore, when the engine speedincreases, the ECU 10 increases the SI ratio according to the speedincrease.

As illustrated in FIG. 11, the ECU 10 of this configuration examplelinearly increases the SI ratio as the engine speed increases. Since thehighest speed N2 is the boundary with the high speed range (D) in whichonly the SI combustion is performed, the SI ratio at the speed N2 isadjusted to 100%. Therefore, the transition between the medium loadrange (B), in which the SPCCI combustion is performed, and the highspeed range (D), in which the SI combustion is performed, is smoothlyperformed without the combustion state greatly varying.

In other words, the ratio of the CI combustion is continuously reducedby adjusting the SI ratio according to the engine speed, and the engineload range in which the combustion by self-ignition is performed isextended as high as possible. As a result, in the engine 1, not only tothe higher engine load side but also to the higher engine speed side,the operating range in which the combustion by self-ignition isperformed may be extended while keeping NVH, which may become an issue,below the allowable value. Therefore, the engine 1 excels in fuelefficiency. Note that such adjustment of the SI ratio is achievable by asophisticated control of the engine 1 executed by the ECU 10 (describedlater in detail).

When the engine load is high, the amount of fuel increases and theamount of heat which the combustion chamber 17 receives also increases,and therefore, the temperature of the combustion chamber 17 becomesrelatively high. Thus, the SI combustion in the SPCCI combustion becomessharper at the high engine load than the at low engine load. Therefore,when comparing at the same engine speed, the SI ratio is adjusted to behigher at a high engine load than at a low engine load. Therefore, asindicated by the one-dotted chain line in FIG. 11, the straight lineindicating the change in the SI ratio with respect to the engine speed(SI ratio/engine speed) has a gentler slope as the engine loadincreases.

(Method of Increasing SI Ratio at High Engine Speed)

At a high engine speed, the combustion cycle is shorter than at a lowengine speed. Since the amount of heat which an inner wall of thecombustion chamber, etc. receives during combustion accordinglydecreases, the temperature of the combustion chamber 17 drops comparedto when the engine speed is low. It also becomes less likely for themixture gas to receive heat from the combustion chamber 17.Additionally, the time length for the mixture gas to combust becomesshort. At the high engine speed, the SI ratio needs to be increasedunder such a disadvantageous condition.

In this regard, by increasing the temperature of the mixture gas itselfin the combustion chamber 17, the combustion is stimulated even underthe condition in which the heat is not easily transferred from thecombustion chamber 17.

Furthermore, the ignition of the CI combustion occurs after the SIcombustion. Although increasing the temperature of the mixture gasstimulates both the SI combustion and the CI combustion, the combustionduration is short when the engine speed is high. Thus, the SI combustionbecomes sharper and is stimulated than the CI combustion, which startslater than the SI combustion. Therefore, by increasing the temperatureof the mixture gas, the SI ratio is increased when the engine speed ishigh.

(First Method)

Therefore, in the engine 1, in order to change the SI ratio according tothe change in the engine speed, the internal EGR ratio is adjustedwithin the medium load range (B) in which the SPCCI combustion isperformed.

For example, the ECU 10 outputs a control signal to the intakeelectrically-operated S-VT 23 and/or the exhaust electrically-operatedS-VT 24, respectively, to change the open timing of the intake valve 21and/or the exhaust valve 22. In this manner, the negative overlap periodis provided, the length of the period is changed, and thehigh-temperature burned gas (internal EGR gas) introduced into thecombustion chamber 17 is set to a given amount. Thus, an internal EGRratio, specifically “a mass ratio (%) of the internal EGR gas withrespect to the total gas within the combustion chamber 17,” is changed.

As illustrated in FIG. 7A, in the engine 1 adopting the operating rangemap 700, the medium load range (B) is divided into the first medium loadsegment (B1) located on the higher engine load side, and the secondmedium load segment (B2) located on the lower engine load side. Withinthe first medium load segment (B1), while corresponding to the highengine load, since the operating range in which the SPCCI combustion isperformed is extended to the higher load side, the boosting isperformed, whereas, within the second medium load segment (B2), theboosting is not performed.

In this engine 1, within the first medium load segment (B1) in which theboosting is performed, since the positive overlap period is provided,even when the internal EGR gas is introduced, it is scavenged by theboosting pressure. Therefore, within the first medium load segment (B1),since it is difficult to change the internal EGR ratio, the first methodis not adopted. Within the second medium load segment (B2), the ECU 10outputs the control signal to the intake electrically-operated S-VT 23etc. so that the internal EGR ratio becomes higher at a high engine loadthan at a low engine load.

FIG. 12A illustrates a relationship between the engine speed and theinternal EGR ratio (internal EGR ratio/engine speed) at given loadswithin the second medium load segment (B2). The ECU 10 linearlyincreases the internal EGR ratio as the engine speed increases from thespeed N1 which is the lowest speed. As a result, the temperature of themixture gas in the combustion chamber 17 rises, and therefore, even whenthe amount of heat received from the combustion chamber 17 is small andthe combustion duration is short, the SI combustion becomes sharp andthe SI ratio increases. Since the ratio of the CI combustion decreasesas the SI ratio increases, the combustion noise caused by the SPCCIcombustion is reduced. When the engine speed is high, NVH is suppressedbelow the allowable value. Note that the relationship between the enginespeed and the internal EGR ratio illustrated in FIGS. 12A and 12B areexamples, and the change rate is suitably changeable according to thespecifications.

As illustrated in FIG. 12A, when the engine speed exceeds a given speed(limitation starting speed Na), the increase in the internal EGR ratiois limited. For example, in order to change the internal EGR ratioaccording to the change in the engine speed, the memory 102 of the ECU10 stores a given map, and the limitation starting speed Na is set inthe map. When the engine speed exceeds the limitation starting speed Na,the ECU 10 outputs the control signal to the intakeelectrically-operated S-VT 23, etc. so as to limit the increase in theinternal EGR ratio.

Note that the limitation starting speed Na is at least set higher than amiddle value of the range of the engine speed (between the speed N1 andthe speed N2) within the medium load range (B). Particularly, when therange of the engine speed within the medium load range (B) is evenlydivided into three segments, the speed Na is preferably set within thehigh-speed segment.

In this configuration example, the internal EGR ratio reaches its peakonce exceeding the limitation starting speed Na and is keptsubstantially constant (depending on the specifications, the value mayslightly vary). If the internal EGR ratio increases and the temperatureof the mixture gas excessively rises, the CI combustion also becomessharp and the combustion noise increases, which may cause NVH above theallowable value. Therefore, the ECU 10 sets the limitation startingspeed Na lower than the value which may cause such a situation, so as tosuppress NVH below the allowable value.

When the engine load is high, the fuel amount increases and the heatreception amount of the combustion chamber 17 also increases, andtherefore, the temperature of the combustion chamber 17 relativelyrises. Thus, the temperature of the mixture gas becomes higher and theSI combustion in the SPCCI combustion becomes sharp at a high engineload than at a low engine load. For this reason, the ECU 10 changes theinternal EGR ratio to be lower over the entire engine speed range as theengine load increases.

Further here, the limitation starting speed Na is set lower as theengine load increases. Since the temperature of the mixture gas tends toincrease excessively as the engine load increases, by setting thelimitation starting speed Na lower as the engine load increases, NVH ismore reliably prevented from exceeding the allowable value.

FIG. 12A illustrates a relationship between the engine speed and theinternal EGR ratio in the case of introducing the internal EGR gas bysetting the negative overlap period (also referred to as “NVO setting”).FIG. 12B illustrates a relationship between the engine speed and theinternal EGR ratio in the case of introducing the internal EGR gas bysetting the positive overlap period (also referred to as “PVO setting”).

In the introduction of the internal EGR gas by the PVO setting, unlikethe introduction of the internal EGR gas by the NVO setting, whichconfines the high-temperature burned gas as it is in the combustionchamber 17, the high-temperature burned gas flows out to the intake port18 once and is introduced into the combustion chamber 17 again. Thehigh-temperature burned gas introduced into the combustion chamber 17,that is, the internal EGR gas, is cooled by flowing out to the intakeport 18 where the temperature is low.

Therefore, when the internal EGR gas is introduced by the PVO setting,compared to the introduction of the internal EGR gas by the NVO setting,the temperature of the combustion chamber 17 becomes relatively lower.Thus, even if the internal EGR ratio increases, the temperature of themixture gas does not easily rise excessively. As a result, the settingof the limitation starting speed Na becomes unnecessary or even if it isneeded, it is set sufficiently higher than when the NVO is set.

Therefore, in the introduction of the internal EGR gas by the PVOsetting, the internal EGR ratio continuously increases without anylimitation until reaching the highest speed N2. Although the limitationstarting speed Na is set close to the highest speed N2 at a high engineload illustrated in FIG. 12B, the setting of the limitation startingspeed Na may be unnecessary even at the high engine load. There may alsobe a case where the setting of the limitation starting speed Na isnecessary at a low engine load or a medium engine load. Note that alsoin this case, the limitation starting speed Na is set closer to thehighest speed N2 compared to the introduction of the internal EGR gas bythe NVO setting.

Note that in the engine 1 adopting one of an operating range map 701 andan operating range map 702 described later, a swirl flow having astrength higher than a given strength is formed in the combustionchamber 17. With such a swirl flow formed in the combustion chamber 17,the mixture gas in the combustion chamber 17 is stratified by the flow.Thus, the mixture gas near the spark plug 25 is made suitable for the SIcombustion, and as a result, the stable SPCCI combustion is achieved.

Therefore, in this case, even if the internal EGR ratio is increased ata high engine speed in order to secure the high SI ratio, by stratifyingthe mixture gas, the EGR gas near the spark plug 25 is reduced. Thus,the stability of the ignition in the SI combustion is secured and stableSPCCI combustion is achieved.

(Second Method)

By the first method alone, in the medium load range (B) in which theSPCCI combustion is performed, the operating range in which thecombustion is by self-ignition is extended to the higher engine speedside only for the second medium load segment (B2). Therefore, in thisengine 1, also within the first medium load segment (B1), similar to thesecond medium load segment (B2), the second method is adopted so thatthe temperature of the mixture gas itself in the combustion chamber 17is changed higher as the engine load increases.

That is, this engine 1 changes the external EGR ratio within the mediumload range (B) in which the SPCCI combustion is performed, so as tochange the SI ratio according to the change in the engine speed.

For example, the introduction amount of the cooled burned gas (externalEGR gas) at a lower temperature than the internal EGR gas is changed bychanging the opening of the EGR valve 54, and thus, the external EGRratio, specifically “a mass ratio (%) of the external EGR gas withrespect to the total gas within the combustion chamber 17” is changed.

The ECU 10 outputs the control signal to the EGR valve 54 so that theexternal EGR ratio decreases lower at a low engine speed than a highengine speed. The external EGR gas has a lower temperature than themixture gas in the combustion chamber 17. Thus, when the introductionamount of the external EGR gas is large, the temperature of the mixturegas drops accordingly. Therefore, by reducing the external EGR ratio,the temperature of the mixture gas itself in the combustion chamber 17is changed higher.

As illustrated in FIG. 7A, the external EGR gas is introduced withinboth the first medium load segment (B1) and the second medium loadsegment (B2), and the amount of the external EGR gas is largeparticularly within the first medium load segment (B1). Therefore, thesecond method is effective within the first medium load segment (B1),and by combining the first method, the SI ratio is effectively adjustedto be high according to the change in the engine speed over the entiremedium load range (B).

FIG. 13A illustrates a relationship between the engine speed and theexternal EGR ratio (external EGR ratio/engine speed) at given loadswithin the medium load range (B). The solid line indicates therelationship within the first medium load segment (B1), and theone-dotted dashed line indicates the relationship within the secondmedium load segment (B2).

As described above, in the engine 1 adopting the operating range map700, the SCV 56 is substantially fully opened in each operating rangeincluding the medium load range (B) in which the SPCCI combustion isperformed. FIG. 13A illustrates such a relationship as described abovein the case where substantially no swirl flow is generated in thecombustion chamber 17 or the flow is weak even if it is generated.

The ECU 10 keeps the external EGR ratio substantially fixed until theengine speed reaches a given speed (decrease starting speed Nb) from thelowest speed N1. In this manner, a high EGR ratio (external EGR ratio)is set at a low engine speed, therefore, even if the value of G/F of themixture gas which needs to be secured is high, it may be settled withinan appropriate range. Since the dilution ratio of the mixture gas iskept high, the fuel efficiency improves.

Then, once exceeding the decrease starting speed Nb, the ECU 10 reducesthe external EGR ratio according to the engine load. For example, inorder to change the external EGR ratio according to the change in theengine speed, the memory 102 of the ECU 10 stores a given map, and inthe map, the decrease starting speed Nb is set in advance.

The decrease starting speed Nb also, similarly to the limitationstarting speed Na, is at least set higher than a middle value of therange of the engine speed within the medium load range (B).Particularly, when the range of the engine speed within the medium loadrange (B) is evenly divided into three segments, the decrease startingspeed Nb is preferably set the same as the limitation starting speed Na.

Based on the map, the ECU 10 outputs the control signal to the EGR valve54 so that the external EGR ratio starts to decrease according to theengine load once exceeding the decrease starting speed Nb. Note thatwhen the engine load is low, the value of the external EGR ratio may bekept as it is even when exceeded the decrease starting speed Nb.

Thus, in a higher speed range at or above the decrease starting speedNb, the ratio of the external EGR gas introduced into the combustionchamber 17 decreases, and the temperature of the mixture gas itselfrises. Therefore, even if the heat reception amount from the combustionchamber 17 is small and the combustion duration is shortened, the SIcombustion becomes sharp and the SI ratio increases. Since the CIcombustion decreases as the SI ratio increases, the combustion noisegenerated by the SPCCI combustion is reduced. When the engine speed ishigh, NVH is suppressed below the allowable value. Note that therelationship between the engine speed and the external EGR ratioillustrated in FIGS. 13A and 13B are examples, and the change ratethereof is suitably changeable according to the specifications.

Here, the ECU 10 outputs the control signal to the EGR valve 54 so thatthe external EGR ratio decreases sharper (the decrease rate increases)as the engine load increases.

When the engine load increases, the fuel amount increases accordingly.If the amount of burned gas is excessively large with respect to theincreasing fuel, the combustion by spark-ignition, that is, the SIcombustion in the SPCCI combustion, becomes unstable and the SI ratiomay decrease. Therefore, the decrease rate of the external EGR ratio isincreased as the engine load increases to stabilize the ignition, a highSI ratio is maintained.

The ECU 10 changes the external EGR ratio also within the second mediumload segment (B2) similarly to the first medium load segment (B1). Inthis configuration example, the EGR gas introduced within the firstmedium load segment (B1) only contains the external EGR gas, while theinternal EGR gas is also introduced within the second medium loadsegment (B2). Therefore, within the second medium load segment (B2), theexternal EGR ratio is changed in consideration of the state functioninside the combustion chamber 17, such as the G/F of the mixture gas dueto the relationship with the introduction amount of the internal EGRgas. As a result, the external EGR ratio within the second medium loadsegment (B2) decreases relatively lower than and completely differentfrom that within the first medium load segment (B1).

In the engine 1 adopting one of the operating range map 701 and theoperating range map 702 described later, different from the engine 1adopting the operating range map 700, a swirl flow stronger than a givenstrength is formed in the combustion chamber 17. FIG. 13B illustratesthe relationship described above in such an engine 1.

By forming the swirl flow stronger than the given strength inside thecombustion chamber 17, the fluidity of the mixture gas in the combustionchamber 17 is enhanced. Thus, even when the amount of EGR gas is large,the amount of EGR gas around the spark plug 25 is reduced when theignition is performed. Therefore, even if the engine load increases andthe fuel amount increases, the stability of the combustion by ignitionis secured, and therefore, the SI combustion in the SPCCI combustion isstably performed.

Thus, as illustrated in FIG. 13B, the ECU 10 outputs the control signalto the EGR valve 54 so that the decrease rate of the external EGR ratiois kept substantially constant with respect to the change in the engineload. Even when the engine load increases, the ECU 10 reduces theexternal EGR ratio in the same state as when the engine load is low.

(Third Method)

In the method of increasing the temperature of the mixture gas as in thefirst and second methods, if the temperature rises excessively, the CIcombustion becomes sharp and the combustion noise increases, which maycause NVH to exceed the allowable value. Further, since the internal EGRratio and the external EGR ratio are restricted by other state function,such as the G/F of the mixture gas, simply changing the internal EGRratio and the external EGR ratio limits the effect of reducing thecombustion noise in the SPCCI combustion, and the effect may not beobtained sufficiently.

Particularly in the first method, the increase in the internal EGR ratiois limited within the range above the limitation starting speed Na. Alsoin the second method, if the external EGR ratio is excessively reducedwithin the range above the decrease starting speed Nb, the temperatureof the mixture gas becomes excessively high and the combustion noiseincreases, which may cause NVH to exceed the allowable value, andtherefore, the decrease rate is limited. Therefore, within the rangesabove the limitation starting speed Na and the decrease starting speedNb, the SI ratio may not increase to a desired SI ratio. In this case,the effect of reducing the combustion noise is limited, and a sufficienteffect cannot be obtained.

Therefore, in this engine 1, another method different from the method ofincreasing the temperature of the mixture gas is used so that the SIratio is also increased thereby.

That is, in conjunction with the first or second method, the ECU 10 usesa third method in which the SCV 56 is controlled to strengthen the flowof the intake air. If the intake air flow is strengthened, the SI ratioincreases because the SI combustion becomes sharp. As a result, thecombustion noise of the SPCCI combustion is reduced.

FIG. 14A illustrates a relationship between the speed of the engine 1adopting the operating range map 700 and the SCV 56. As described above,in the engine 1 adopting the operating range map 700, the SCV 56 issubstantially fully opened in each operating range including the mediumload range (B) in which the SPCCI combustion is performed. FIG. 14Aillustrates the above relationship in such a case. On the other hand,FIG. 14B illustrates the above relationship of the engine 1 adopting anoperating range map 701 of FIG. 7B and the engine 1 adopting anoperating range map 702 of FIG. 7C.

As illustrated in FIG. 14A, in the engine 1 adopting the operating rangemap 700, when the engine speed reaches a given speed (operation startingspeed Nc) at which the adjustment to the desired SI ratio becomesdifficult by the first or second method alone, such as the limitationstarting speed Na at which the limitation of the internal EGR ratio isstarted, the decrease starting speed Nb at which the decrease of theexternal EGR ratio is started, the ECU 10 outputs the control signal tothe SCV 56 to narrow its opening from the substantially full openedstate. Thus, the swirl flow in the combustion chamber 17 becomes strong.The ECU 10 linearly reduces the opening of the SCV 56 as the enginespeed increases. As the engine speed increases, since the swirl flowbecomes stronger, the SI combustion becomes even sharper.

As a result, in the first or second method, even within the higherengine speed range where the limitation is applied, the combustion noiseof the SPCCI combustion is reduced. As a result, when the engine speedis high, NVH is suppressed below the allowable value. By combining thefirst or second method with the third method, the SI ratio is changedhigher and more suitably.

On the other hand, as illustrated in FIG. 14B, in the engine 1 adoptingthe operating range map 701 and the engine 1 adopting the operatingrange map 702 described later, the swirl flow stronger than the givenstrength is formed in the combustion chamber 17 within the range inwhich the SPCCI combustion is performed.

For example, at the timing of reaching the ignition from the intakestroke, the SCV 56 is controlled at a given narrow opening without beingfully closed so that the swirl ratio becomes a range between 1.5 to 3.The opening of the SCV 56 or the swirl ratio is suitably set accordingto the specifications of the engine 1. FIG. 14B illustrates the casewhere the SCV 56 is kept substantially constant at approximately 30%with respect to the engine speed.

When the SPCCI combustion is performed, by forming such a swirl flow inthe combustion chamber 17, the fluidity in the combustion chamber 17 isenhanced, and stratified mixture gas is formed in the combustion chamber17. Thus, the stable SPCCI combustion is achieved within a wideoperating range of the engine 1. Details will be described later.

(Fourth Method)

In the engine 1 adopting the operating range map 700, the SI ratio isincreased to the higher speed side also by changing the ignition timing.

FIG. 15 illustrates a relationship between the engine speed and theignition timing at a given load within the medium load range (B) inwhich the SPCCI combustion is performed. Similar to the third method,when the engine speed reaches a given speed (advance starting speed Nd)at which the adjustment to the desired SI ratio becomes difficult byonly the first or second method, such as the limitation starting speedNa at which the limitation of the internal EGR ratio is started, thedecrease starting speed Nb at which the decrease of the external EGRratio is started, the ECU 10 outputs the control signal to the sparkplug 25 to advance the ignition timing. Thus, the start of the SIcombustion is advanced and the SI combustion becomes sharp. The ECU 10linearly advances the ignition timing as the engine speed increases. TheSI combustion becomes even sharper as the engine speed increases.

The fourth method may replace the third method. The fourth method may beused together with the third method. By suitably combining the first tofourth methods, the adjustment to the stable SI ratio becomes possibleeven at a high engine speed.

(Changing SI Ratio)

FIG. 16 illustrates a process relating to the operation control of theengine 1 executed by the ECU 10. Based on the detection signals from thesensors SW1 to SW16, the ECU 10 determines the operating state of theengine 1, and also changes the state function inside the combustionchamber 17, the injection amount, the injection timing, and the ignitiontiming so that the combustion in the combustion chamber 17 is of the SIratio corresponding to the operating state.

The ECU 10 further adjusts the SI ratio by the SI ratio changing module102 a when its change is determined to be needed, based on the detectionsignals from the sensors.

First at S1, the ECU reads the detection signals of the sensors SW1 toSW16. Next at S2, the ECU 10 determines the operating state of theengine 1 based on the detection signals and sets the target SI ratiowhich is as illustrated in FIG. 9 or 11.

Next at S3, the ECU 10 sets a target in-cylinder state function forachieving the set target SI ratio, based on a preset combustion model.For example, the target temperature, the target pressure, and the targetstate function in the combustion chamber 17 are set. At S4, the ECU 10sets the opening of the EGR valve 54, the opening of the throttle valve43, the opening of the air bypass valve 48, the phase angles of theintake electrically-operated S-VT 23 and the exhaustelectrically-operated S-VT 24, and the opening of the SCV 56, which arerequired for achieving the target in-cylinder state function.

The ECU 10 sets the control amounts of these devices in advance based ona map stored in the ECU 10. Based on the set control amounts, the ECU 10outputs control signals to the EGR valve 54, the throttle valve 43, theair bypass valve 48, the intake electrically-operated S-VT 23, theexhaust electrically-operated S-VT 24, and the SCV 56. As each deviceoperates based on the control signal from the ECU 10, the state functionin the combustion chamber 17 becomes the target state function.

The ECU 10 further calculates a predicted value and an estimated valueof the state function in the combustion chamber 17 based on the setcontrol amount of each device. The state function predicted value isobtained by predicting the state function in the combustion chamber 17before the intake valve 21 is closed, and is used for setting the fuelinjection amount on the intake stroke as described later. The statefunction estimated value is obtained by estimating the state function inthe combustion chamber 17 after the intake valve 21 is closed, and isused for setting of the fuel injection amount and the ignition timing onthe compression stroke as described later. The state function estimatedvalue is also used for calculating a state function error based oncomparison with an actual combustion state as described later.

At S5, the ECU 10 sets the fuel injection amount on the intake strokebased on the state function predicted value. Note that in a case wherethe fuel injection is not performed on the intake stroke, the fuelinjection amount is zero. At S6, the ECU 10 controls the injection ofthe injector 6. That is, the ECU 10 outputs a control signal to theinjector 6 to inject the fuel into the combustion chamber 17 at a giveninjection timing.

At S7, the ECU 10 sets the fuel injection amount on the compressionstroke based on the state function estimated value and the fuelinjection result on the intake stroke. Note that in the case where thefuel injection is not performed on the compression stroke, the fuelinjection amount is zero. In the case where the split injections areperformed on the compression stroke, the injection amount of thefirst-stage injection and the injection amount of the second-stageinjection are respectively set. At S8, the ECU 10 outputs a controlsignal to the injector 6 to inject the fuel into the combustion chamber17 at the injection timing based on the preset map.

At S9, the ECU 10 sets the ignition timing based on the state functionestimated value and the fuel injection result on the compression stroke.At S10, the ECU 10 outputs a control signal to the spark plug 25 toignite the mixture gas into the combustion chamber 17 at the setignition timing.

The SI combustion or SPCCI combustion is performed in the combustionchamber 17 by the spark plug 25 igniting the mixture gas. At S11, theECU 10 reads a change in pressure inside the combustion chamber 17detected by the pressure sensor SW6, and, based on this change,determines the combustion state of the mixture gas in the combustionchamber 17. At S12, the ECU 10 further compares the detection result ofthe combustion state with the state function estimated value estimatedat S4 to calculate a difference (error) between the state functionestimated value and the actual state function. The calculated error isused in the estimation at S4 in the next cycle and/or its subsequentcycles.

The ECU 10 changes the openings of the throttle valve 43, the EGR valve54, the SCV 56, and/or the air bypass valve 48, and the phase angles ofthe intake electrically-operated S-VT 23 and the exhaustelectrically-operated S-VT 24. Thus, the introduction amounts of freshair and EGR gas amount into the combustion chamber 17 are changed. Thefeedback of the state function error is equivalent to changing the SIratio when the ECU 10 determines that the change of the SI ratio isneeded based on the error between the target SI ratio and the actual SIratio (error).

If, at S8, the temperature in the combustion chamber 17 is estimated tobe lower than the target temperature based on the state functionestimated value, the ECU 10 advances the injection timing on thecompression stroke than the injection timing which is based on the mapso that the ignition timing is advanced. On the other hand, at S8, ifthe temperature in the combustion chamber 17 is estimated to be higherthan the target temperature based on the state function estimated value,the ECU 10 retards the injection timing on the compression stroke thanthe injection timing which is based on the map so that the ignitiontiming retards.

That is, as indicated by P2 of FIG. 17, if the temperature in thecombustion chamber 17 is low, the timing θ_(CI) at which the unburnedmixture gas self-ignites after the SI combustion is started by thespark-ignition retards, and the SI ratio shifts from the target SI ratio(see P1). In this case, the unburned fuel increases and the exhaust gasperformance degrades.

Therefore, when the temperature in the combustion chamber 17 isestimated to be lower than the target temperature, the ECU 10 advancesthe injection timing as well as the ignition timing θ_(IG) at S10 ofFIG. 16. As indicated by P3 of FIG. 17, since the start of SI combustionis advanced, sufficient heat generation by the SI combustion occurs, andtherefore, the timing θ_(CI) of the self-ignition of the unburnedmixture gas is prevented from retarding when the temperature in thecombustion chamber 17 is low. As a result, the SI ratio approaches thetarget SI ratio. The increase of the unburned fuel and the degradationof exhaust gas performance are prevented.

Moreover, as indicated by P4 of FIG. 17, when the temperature in thecombustion chamber 17 is high, the unburned mixture gas self-ignitesimmediately after the SI combustion is started by the spark-ignition,and the SI ratio shifts from the target SI ratio (see P1). In this case,the combustion noise increases.

Therefore, when the temperature in the combustion chamber 17 isestimated to exceed the target temperature, the ECU 10 retards theinjection timing as well as the ignition timing θ_(IG) at S10 of FIG.16. As indicated by P5 of FIG. 17, since the start of SI combustion isretarded, the timing θ_(CI) of the self-ignition of the unburned mixturegas is prevented from advancing when the temperature in the combustionchamber 17 is high. As a result, the SI ratio approaches the target SIratio. The increase of the combustion noise is avoided.

The changes of the injection timing and the ignition timing areequivalent to changing the SI ratio when the ECU 10 determines that theSI ratio in the SPCCI combustion is required to be changed. By changingthe injection timing, suitable mixture gas is formed in the combustionchamber 17 at the advanced or retarded ignition timing. The spark plug25 reliably ignites the mixture gas, and the unburned mixture gasself-ignites at a suitable timing.

Note that the change of the state function inside the combustion chamber17 through controlling the throttle valve 43, the EGR valve 54, the airbypass valve 48, the intake electrically-operated S-VT 23, the exhaustelectrically-operated S-VT 24, and the SCV 56 based on the actualcombustion state in FIG. 17 is as described at S12 and S4 of FIG. 16.

The engine 1 changes the SI ratio by the state function setting devicesincluding the throttle valve 43, the EGR valve 54, the air bypass valve48, the intake electrically-operated S-VT 23, the exhaustelectrically-operated S-VT 24, and the SCV 56. The SI ratio may roughlybe changed by changing the state function inside the combustion chamber17. At the same time, the engine 1 changes the SI ratio by changing theinjection timing and the ignition timing of the fuel. By changing theinjection timing and the ignition timing, for example, the differencebetween the cylinders may be corrected and the self-ignition timing isfinely changed. By changing the SI ratio in two stages, the engine 1accurately achieves the target SPCCI combustion corresponding to theoperating state.

<Second Operating Range Map>

FIG. 7B illustrates a second configuration example of the operatingrange map of the engine 1 (second operating range map 701). Theoperating range map 701 is divided into the following three ranges.

(A): a low load range including the idle operation

(B): a medium load range between the low load range (A) and thefollowing high load range (C)

(C): a high load range including the full engine load

That is, on the operating range map 701, compared to the operating rangemap 700 described above, each of the low load range (A), the medium loadrange (B), and the high load range (C) is extended in the engine speeddirection.

In the engine 1 adopting the operating range map 701, especially inorder to extend the medium load range (B) in which the SPCCI combustionis performed, the swirl flow stronger than a given strength is generatedin the combustion chamber 17.

For example, the ECU 10 adjusts the opening of the SCV 56 as needed.When the operating state of the engine 1 is within the medium load range(B), the SCV 56 is adjusted to be fully closed or have a given narrowopening. As a result, a swirl flow having the given strength is formedin the combustion chamber 17. At the timing of reaching the ignitionfrom the intake stroke, the swirl ratio may be 4 or higher. Moreover,the swirl ratio may be set below 4, for example, within a range of 1.5to 3.

Also in the engine 1 adopting the operating range map 701, the SI ratiois preferably adjusted similar to the engine 1 adopting the operatingrange map 700. That is, as illustrated in FIG. 11, within the secondmedium load segment (B2) in which the boosting is not performed, the SIratio is linearly increased as the engine speed increases. As a result,also within the high speed range, the SPCCI combustion is stablyperformed while keeping NVH, which can become an issue, below theallowable value.

Further in the engine 1 adopting the operating range map 701, the SIratio is increased at a high engine speed by the first to fourth methodssimilar to the engine 1 adopting the operating range map 700.

<Third Operating Range Map>

FIG. 7C illustrates a third configuration example of the operating rangemap of the engine 1 (third operating range map 702).

(1)-1: a low load range including the idle operation and extending inlow and medium engine speed ranges

(1)-2: a medium load range in which the engine load is higher than thelow load range and extending in the low and medium engine load ranges

(2): a high-load medium-speed range in which the engine load is higherthan the medium load range and which is located in the high load rangeincluding the full engine load

(3): a high-load low-speed range located in the high load range and inwhich the engine speed is lower than the high-load medium-speed range

(4): and a high speed range in which the engine speed is higher than thelow load range, the medium load range, the high-load medium-speed range,and the high-load low-speed range

Here, the low speed range, the medium speed range, and the high speedrange may be defined by substantially evenly dividing, in the enginespeed direction, the entire operating range of the engine 1 into threeranges of the low speed range, the medium speed range and the high speedrange. On the operating range map 702, the engine speed lower than aspeed N1 is defined as low, the engine speed higher than a speed N2 isdefined as high, and the engine speed between engine speeds N1 and N2 isdefined as medium.

For example, the speed N1 may be about 1,200 rpm and the speed N2 may beabout 4,000 rpm. Note that the speed N1 and the speed N2 may be the sameor different between the operating range map 700 and the operating rangemap 702.

On the operating range map 702, mainly for improving the fuel efficiencyand the exhaust gas performance, the engine 1 performs the SPCCIcombustion within the low load range (1)-1, the medium load range (1)-2,and the high-load medium-speed range (2). The operating range map 702 isdifferent from the operating range map 701 in that the SPCCI combustionis performed also when the engine 1 operates at a low load and when theengine 1 operates at a high load. Further, the engine 1 performs thecombustion by spark-ignition within the other ranges, specifically, thehigh-load low-speed range (3) and the high speed range (4). Hereinafter,the operation of the engine 1 within each range is described in detailwith reference to the fuel injection timing and the ignition timingillustrated in FIG. 18.

(Low Load Range (1)-1)

When the engine 1 is operating within the low load range (1)-1, theengine 1 performs the SPCCI combustion.

The reference character 601 in FIG. 18 indicates one example of the fuelinjection timings (reference characters 6011 and 6012) and the ignitiontiming (reference character 6013), and a combustion waveform (i.e., awaveform indicating a change in heat generation rate with respect to thecrank angle; the reference character 6014) when the engine 1 isoperating in the operating state of the reference character 601 withinthe low load range (1)-1 of the engine 1.

When the engine 1 is operating within the low load range (1)-1, theswirl flow having the given strength is formed in the combustion chamber17 (see FIG. 14B). The swirl ratio when the engine 1 operates within thelow load range (1)-1 is approximately 2, and is adjusted within a rangeof below 4, for example, within the range of 1.5 to 3. Note that astrong swirl flow may be formed by setting the swirl ratio to 4 orhigher.

The swirl flow is strong in an outer circumferential portion of thecombustion chamber 17 and weak in a center portion. The SCV 56 iscontrolled to have a given narrow opening without being fully closed. Asdescribed above, since the intake port 18 is a tumble port, an obliqueswirl flow having a tumble component and a swirl component is formed inthe combustion chamber 17.

When the engine 1 operates within the low load range (1)-1, the air-fuelratio (A/F) of the mixture gas is leaner than the theoretical air-fuelratio in the entire combustion chamber 17. That is, the excess air ratioλ of the mixture gas exceeds 1 in the entire combustion chamber 17. Morespecifically, the A/F of the mixture gas in the entire combustionchamber 17 is 30:1 or higher. In this manner, generation of raw NO_(x)is reduced and the exhaust gas performance is improved.

When the engine 1 operates within the low load range (1)-1, the EGRsystem 55 introduces the EGR gas into the combustion chamber 17 asneeded. For example, the internal EGR gas is introduced according to thesetting of the positive overlap period.

When the engine 1 operates within the low load range (1)-1, the mixturegas is stratified between the center portion and the outercircumferential portion of the combustion chamber 17. The center portionof the combustion chamber 17 is the portion where the spark plug 25 isdisposed, and the outer circumferential portion is the portion aroundthe center portion and in contact with a liner of the cylinder 11. Thecenter portion of the combustion chamber 17 may be defined as a portionwhere the swirl flow is weak and the outer circumferential portion maybe defined as a portion where the swirl flow is strong.

The fuel concentration of the mixture gas in the center portion ishigher than that in the outer circumferential portion. For example, theA/F of the mixture gas in the center portion is between 20:1 and 30:1,and the A/F of the mixture gas in the outer circumferential portion is35:1 or higher.

When the engine 1 operates within the low load range (1)-1, the injector6 basically injects the fuel into the combustion chamber 17 by splittingit into in a plurality of injections on the compression stroke. Themixture gas is stratified in the center portion and the outercircumferential portion of the combustion chamber 17 by the splitinjections of the fuel and the swirl flow in the combustion chamber 17.

After the fuel injection is ended, the spark plug 25 ignites the mixturegas in the center portion of the combustion chamber 17 at the giventiming before CTDC (see the reference character 6013). Since the fuelconcentration of the mixture gas in the center portion is relativelyhigh, ignitability improves and the SI combustion by the flamepropagation stabilizes. By stabilizing the SI combustion, the CIcombustion starts at the suitable timing. The controllability of the CIcombustion improves in the SPCCI combustion. As a result, when theengine 1 operates within the low load range (1)-1, both the reduction ofthe generation of combustion noise and the improvement of the fuelefficiency by shortening the combustion period are achieved.

(Medium Load Range (1)-2)

Also when the engine 1 is operating within the medium load range (1)-2,the engine 1 performs the SPCCI combustion similarly to the low loadrange (1)-1. The medium load range (1)-2 corresponds to the medium loadrange (B) on the operating range map 701.

The reference character 602 in FIG. 18 indicates one example of the fuelinjection timings (reference characters 6021 and 6022) and the ignitiontiming (reference character 6023), and a combustion waveform (referencecharacter 6024) when the engine 1 is operating in the operating state ofthe reference character 602 within the medium load range (1)-2 of theengine 1.

The EGR system 55 introduces the EGR gas into the combustion chamber 17when the operating state of the engine 1 is within the medium load range(1)-2. For example, the internal EGR gas is introduced according to thesetting of the positive overlap period.

Also when the engine 1 operates within the medium load range (1)-2,similar to the low load range (1)-1, a swirl flow having a givenstrength, in which the swirl ratio is about 2, is formed in thecombustion chamber 17. The SCV 56 is not fully closed but has a givennarrow opening. By forming the swirl flow, the residual gas accumulatedin the cavity 31 is led outside the cavity 31. As a result, the G/F ofthe mixture gas in the SI portion near the spark plug 25 may be variedfrom the G/F of the mixture gas in the CI portion around the SI portion.Therefore, the SPCCI combustion is stabilized by setting the total G/Fin the entire combustion chamber 17 to between 18 and 50 as describedabove.

Further, since the turbulence kinetic energy in the combustion chamber17 increases by forming the swirl flow, when the engine 1 operateswithin the medium load range (1)-2, the flame of the SI combustionpropagates promptly and the SI combustion is stabilized. Thecontrollability of the CI combustion improves by stabilizing the SIcombustion. By making the timing of the CI combustion in the SPCCIcombustion appropriate, the generation of combustion noise is reducedand the fuel efficiency is improved. Further, the variation in torquebetween cycles is reduced.

When the engine 1 operates within the medium load range (1)-2, theair-fuel ratio (A/F) of the mixture gas is at the theoretical air-fuelratio (A/F≈14.7:1) in the entire combustion chamber 17. The three-waycatalyst purifies the exhaust gas discharged from the combustion chamber17. Thus, the exhaust gas performance of the engine 1 becomes good. TheA/F of the mixture gas may be set to remain within the purificationwindow of the three-way catalyst. Therefore, the excess air ratio λ ofthe mixture gas may be set to 1.0±0.2.

When the engine 1 operates within the medium load range (1)-2, theinjector 6 performs the fuel injection on the intake stroke (referencecharacter 6021) and the fuel injection on the compression stroke(reference character 6022). By performing the first injection 6021 onthe intake stroke, the fuel is distributed substantially evenly in thecombustion chamber 17. By performing the second injection 6022 on thecompression stroke, the temperature in the combustion chamber 17 isreduced by latent heat of vaporization of the fuel. The mixture gascontaining the fuel injected in the first injection 6021 is preventedfrom causing pre-ignition.

When the injector 6 performs the first injection 6021 on the intakestroke and the second injection 6022 on the compression stroke, themixture gas with the excess air ratio λ of 1.0±0.2 as a whole is formedin the combustion chamber 17. Since the fuel concentration of themixture gas is substantially homogeneous, the improvement in the fuelefficiency by reducing the unburned fuel loss and the improvement in theexhaust gas performance by avoiding the smoke generation are achieved.The excess air ratio λ is preferably 1.0 to 1.2. Further, the total G/Fin the entire combustion chamber 17 is 18 to 50, and the G/F of the SIportion near the spark plug 25 is 14 to 22.

By the spark plug 25 igniting the mixture gas at the given timing beforeCTDC (reference character 6023), the mixture gas combusts by flamepropagation. After the combustion by flame propagation is started, theunburned mixture gas self-ignites at the target timing and causes the CIcombustion. The fuel injected in the second-stage injection mainlycauses the SI combustion. The fuel injected in the first-stage injectionmainly causes the CI combustion. By having the total G/F in the entirecombustion chamber 17 between 18 and 50 and the G/F of the SI portionnear the spark plug 25 between 14 and 22, the SPCCI combustion isstabilized.

Here, as illustrated in FIG. 7C, a range in which the booster 44 isturned off (refer to S/C OFF) is a part of the low load range (1)-1 anda part of the medium load range (1)-2. In detail, the booster 44 isturned off within the low engine speed segment of the low load range(1)-1. Within a high engine speed segment of the low load range (1)-1,the booster 44 is turned on to increase the boosting pressure in orderto secure a required intake charge amount corresponding to the enginespeed being high. Further, within a low-load low-speed segment of themedium load range (1)-2, the booster 44 is turned off. Within a highload segment of the medium load range (1)-2, the booster 44 is turned onin order to secure the required intake air charge amount correspondingto the engine speed being high. Within a high speed segment of themedium load range (1)-2, the booster 44 is turned on in order to securethe required intake air charge amount corresponding to the engine speedbeing high.

Note that within the high-load medium-speed range (2), the high-loadlow-speed range (3), and the high speed range (4), the booster 44 isturned on throughout the ranges.

(High-Load Medium-Speed Range (2))

Also when the engine 1 is operating within the high-load medium-speedrange (2), the engine 1 performs the SPCCI combustion similarly to thelow load range (1)-1 and the medium load range (1)-2.

The reference character 603 in FIG. 18 indicates one example of the fuelinjection timings (reference characters 6031 and 6032) and the ignitiontiming (reference character 6033), and a combustion waveform (referencecharacter 6034) when the engine 1 is operating in the operating state ofthe reference character 603 within the high-load medium-speed range (2)of the engine 1. Further, the reference character 604 in FIG. 18indicates one example of the fuel injection timings (referencecharacters 6041) and the ignition timing (reference character 6042), anda combustion waveform (reference character 6043) when the engine speedis higher than in the operating state of the reference character 603.

The EGR system 55 introduces the EGR gas (e.g., external EGR gas) intothe combustion chamber 17 when the operating state of the engine 1 iswithin the high-load medium-speed range (2). The engine 1 reduces theEGR gas amount as the engine load increases. At the full engine load,the EGR gas may be set to zero.

Also when the engine 1 operates within the high-load medium-speed range(2), similar to the low load range (1)-1, a swirl flow having a givenstrength, in which the swirl ratio is below 4, is formed in thecombustion chamber 17. A strong swirl flow may be formed by setting theswirl ratio to 4 or higher. The SCV 56 is controlled to have a givennarrow opening without being fully closed.

When the engine 1 operates within the high-load medium-speed range (2),the air-fuel ratio (A/F) of the mixture gas is at or richer than thetheoretical air-fuel ratio in the entire combustion chamber 17 (i.e.,the excess air ratio λ of the mixture gas is λ≤1).

When the engine 1 operates at the low speed side of the high-loadmedium-speed range (2), the injector 6 injects the fuel on the intakestroke (reference character 6031) and injects the fuel at a final stageof the compression stroke (reference character 6032). The final stage ofthe compression stroke may be defined by evenly dividing the compressionstroke into three stages of an initial stage, an intermediate stage, andthe final stage.

The first-stage injection 6031 starting on the intake stroke may startthe fuel injection in an early half of the intake stroke. The early halfof the intake stroke may be defined by evenly dividing the intake strokeinto two parts of the early half and the latter half. For example, thefirst-stage injection may start the fuel injection at 280° CA beforeTDC.

When the injection of the first-stage injection 6031 is started in theearly half of the intake stroke, although not illustrated, the fuelspray hits an opening edge of the cavity 31 so that a portion of thefuel enters the squish area 171 of the combustion chamber 17 and therest of the fuel enters the section within the cavity 31. The swirl flowis strong in the outer circumferential portion of the combustion chamber17 and weak in the center portion. Therefore, the portion of the fuelthat enters the squish area 171 joins the swirl flow, and the rest ofthe fuel entered the section within the cavity 31 joins the inner sideof the swirl flow. The fuel joins the swirl flow, remains in the swirlflow during the intake stroke and the compression stroke and forms themixture gas for the CI combustion in the outer circumference portion ofthe combustion chamber 17. The fuel that enters the inner side of theswirl flow also remains at the inner side of the swirl flow during theintake stroke and the compression stroke and forms the mixture gas forthe SI combustion in the center portion of the combustion chamber 17.

When the engine 1 operates within the high-load medium-speed range (2),the excess air ratio λ of the mixture gas in the center portion wherethe spark plug 25 is disposed preferably is 1 or less, and the excessair ratio λ of the mixture gas in the outer circumferential portion is 1or less, preferably below 1. The air-fuel ratio (A/F) of the mixture gasin the center portion may be, for example, between 13 and thetheoretical air-fuel ratio (14.7:1). The air-fuel ratio of the mixturegas in the center portion may be leaner than the theoretical air-fuelratio. Further, the air-fuel ratio of the mixture gas in the outercircumferential portion may be, for example, between 11:1 and thetheoretical air-fuel ratio, preferably between 11:1 and 12:1. Since theamount of fuel within the mixture gas increases in the outercircumferential portion when the excess air ratio λ of the outercircumferential portion of the combustion chamber 17 is set to below 1,the temperature is lowered by the latent heat of vaporization of thefuel. The air-fuel ratio of the mixture gas in the entire combustionchamber 17 may be between 12.5:1 and the theoretical air-fuel ratio,preferably between 12.5:1 and 13:1.

The second-stage injection 6032 performed in the final stage of thecompression stroke may start the fuel injection at 10° CA before TDC. Byperforming the second-stage injection immediately before TDC, thetemperature in the combustion chamber 17 is lowered by the latent heatof vaporization of the fuel. Although a low-temperature oxidationreaction of the fuel injected by the first-stage injection 6031progresses on the compression stroke and transitions to ahigh-temperature oxidation reaction before TDC, by performing thesecond-stage injection 6032 immediately before TDC so as to lower thetemperature inside the combustion chamber, the transition from thelow-temperature oxidation reaction to the high-temperature oxidationreaction is avoided and pre-ignition is prevented. Note that the ratiobetween the injection amount of the first-stage injection and theinjection amount of the second-stage injection may be, for example,95:5.

The spark plug 25 ignites the mixture gas in the center portion ofcombustion chamber 17 near CTDC (reference character 6033). The sparkplug 25 ignites, for example, after CTDC. Since the spark plug 25 isdisposed in the center portion of the combustion chamber 17, theignition of the spark plug 25 causes the mixture gas in the centerportion to start the SI combustion by flame propagation.

Within the high load range, the fuel injection amount increases as wellas the temperature of the combustion chamber 17, therefore the CIcombustion is likely to start early. In other words, within the highload range, the pre-ignition of the mixture gas is likely to occur.However, since the temperature of the outer circumferential portion ofthe combustion chamber 17 is lowered by the latent heat of vaporizationof the fuel as described above, the CI combustion is avoided fromstarting immediately after the mixture gas is spark-ignited.

As described above, when the spark plug 25 ignites the mixture gas inthe center portion, the combustion rate increases and the SI combustionis stabilized by high turbulence kinetic energy, and the flame of the SIcombustion propagates in the circumferential direction along the swirlflow inside the combustion chamber 17. Then, at a given position of theouter circumferential portion of the combustion chamber 17 in thecircumferential direction, the unburned mixture gas self-ignites and theCI combustion starts.

In the concept of this SPCCI combustion, by the combination ofstratifying the mixture gas in the combustion chamber 17 and causing theswirl flow in the combustion chamber 17, the SI combustion issufficiently performed until the CI combustion starts. As a result, thegeneration of combustion noise is reduced and, since the combustiontemperature does not become excessively high, generation of NO_(x) isalso reduced. Further, the variation in torque between cycles isreduced.

Further, since the temperature in the outer circumferential portion islow, the CI combustion becomes slower and the generation of thecombustion noise is reduced. Moreover, since the combustion period isshortened by the CI combustion, within the high load range, the torqueimproves and also the thermal efficiency improves. Thus, by performingthe SPCCI combustion within the high engine load range, the engine 1 isimproved in the fuel efficiency while avoiding the combustion noise.

When the engine 1 operates at the high speed side of the high-loadmedium-speed range (2), the injector 6 starts the fuel injection on theintake stroke (reference character 6041).

The first-stage injection 6041 starting on the intake stroke may startthe fuel injection in the early half of the intake stroke similarly tothe above description. For example, the first-stage injection 6041 maystart the fuel injection at 280° CA before TDC. The first-stageinjection may last over the intake stroke and end on the compressionstroke. By setting the start of injection of the first-stage injection6041 in the early half of the intake stroke, the mixture gas for the CIcombustion is formed in the outer circumferential portion of thecombustion chamber 17 and the mixture gas for the SI combustion isformed in the center portion of the combustion chamber 17. Similar tothe above description, the excess air ratio λ of the mixture gas in thecenter portion where the spark plug 25 is disposed preferably is 1 orless, and the excess air ratio λ of the mixture gas in the outercircumferential portion is 1 or less, preferably below 1. The air-fuelratio (A/F) of the mixture gas in the center portion may be, forexample, between 13 and the theoretical air-fuel ratio (14.7:1). Theair-fuel ratio of the mixture gas in the center portion may be leanerthan the theoretical air-fuel ratio. Further, the air-fuel ratio of themixture gas in the outer circumferential portion may be, for example,between 11:1 and the theoretical air-fuel ratio, preferably between 11:1and 12:1. The air-fuel ratio of the mixture gas in the entire combustionchamber 17 may be between 12.5:1 and the theoretical air-fuel ratio,preferably between 12.5:1 and 13:1.

When the engine speed increases, the time length during which the fuelinjected in the first-stage injection 6041 reacts becomes shorter.Therefore, the second-stage injection for suppressing the oxidationreaction of the mixture gas may be omitted.

The spark plug 25 ignites the mixture gas in the center portion ofcombustion chamber 17 near CTDC (reference character 6042). The sparkplug 25 ignites, for example, after CTDC.

As described above, by stratifying the mixture gas, within the high-loadmedium-speed range (2), the combustion noise is reduced and the SPCCIcombustion is stabilized.

(High-Load Low-Speed Range (3))

When the engine 1 is operating within the high-load low-speed range (3),the engine 1 performs the SI combustion instead of the SPCCI combustion.The high-load low-speed range (3) corresponds to the first high loadsegment (C1) on the operating range map 701.

The reference character 605 in FIG. 18 indicates one example of the fuelinjection timings (reference characters 6051 and 6052) and the ignitiontiming (reference character 6053), and a combustion waveform (referencecharacter 6054) when the engine 1 is operating in the operating state ofthe reference character 605 within the high-load low-speed range (3) ofthe engine 1.

The EGR system 55 introduces the EGR gas into the combustion chamber 17when the operating state of the engine 1 is within the high-loadlow-speed range (3). The engine 1 reduces the EGR gas amount as theengine load increases. At the full load, the EGR gas may be set to zero.

When the engine 1 operates within the high-load low-speed range (3), theair-fuel ratio (A/F) of the mixture gas is at the theoretical air-fuelratio (A/F=44.7:1) in the entire combustion chamber 17. The A/F of themixture gas may be set to remain within the purification window of thethree-way catalyst. Therefore, the excess air ratio λ of the mixture gasmay be set to 1.0±0.2. By setting the air-fuel ratio of the mixture gasto the theoretical air-fuel ratio, the fuel efficiency improves withinthe high-load low-speed range (3). When the engine 1 operates within thehigh-load medium-speed range (3), the fuel concentration of the mixturegas in the entire combustion chamber 17 may be set so that the excessair ratio λ is 1 or less and equal to or higher than the excess airratio λ within the high-load medium-speed range (2), preferably higherthan the excess air ratio λ within the high-load medium-speed range (2).

On the operating range map 702, when the engine 1 operates within thehigh-load low-speed range (3), the injector 6 injects the fuel into thecombustion chamber 17 at the timings of the intake stroke and in theretard period from the final stage of the compression stroke to theearly stage of the expansion stroke (reference characters 6051 and6052). By injecting the fuel in two injections, the amount of fuelinjected in the retard period is reduced. By injecting the fuel on theintake stroke (reference character 6051), the formation period of timeof the mixture gas is sufficiently secured. Additionally, by injectingthe fuel in the retard period (reference character 6052), the flow inthe combustion chamber 17 immediately before the ignition isstrengthened, which is advantageous in stabilizing the SI combustion.

After the fuel is injected, the spark plug 25 ignites the mixture gas ata timing near CTDC (reference character 6053). The spark plug 25ignites, for example, after CTDC. The mixture gas causes the SIcombustion on the expansion stroke. Since the SI combustion starts onthe expansion stroke, the CI combustion does not start.

When the engine 1 operates within the high-load low-speed range (3), theswirl flow is made weaker than when operating within the high-loadmedium-speed range (2). When the engine 1 operates within the high-loadlow-speed range (3), the opening of the SCV 56 is larger than whenoperating within the high-load medium-speed range (2). The opening ofthe SCV 56 may be, for example, about 50% (i.e., half opened).

The axes of the nozzle ports of the injector 6 do not circumferentiallyoverlap with the spark plug 25. The fuel injected from the nozzle portsflows in the circumferential direction due to the swirl flow in thecombustion chamber 17. By the swirl flow, the fuel is promptly conveyedto near the spark plug 25. The fuel is vaporized while being conveyed tonear the spark plug 25.

On the other hand, if the swirl flow is excessively strong, the fuelflows in the circumferential direction and reaches away from the sparkplug 25, and the fuel cannot promptly be conveyed to near the spark plug25. For this reason, when the engine 1 operates within the high-loadlow-speed range (3), the swirl flow is made weaker than when operatingwithin the high-load medium-speed range (2). As a result, the fuel ispromptly conveyed to near the spark plug 25, thus the ignitability ofthe mixture gas improves and the SI combustion stabilizes.

(High Speed Range (4))

When the engine speed is high, the time length for the crank angle tochange 1° becomes shorter. Therefore, for example, within the high speedside of the high load range, it is difficult to stratify the mixture gasin the combustion chamber 17 as described above. When the engine speedincreases, it becomes difficult to perform the SPCCI combustiondescribed above.

When the engine 1 is operating within the high speed range (4), theengine 1 performs the SI combustion instead of the SPCCI combustion.Note that the high speed range (4) extends over the entire loaddirection from low load to high loads.

The reference character 606 in FIG. 18 indicates one example of the fuelinjection timing (reference characters 6061) and the ignition timing(reference character 6062), and a combustion waveform (referencecharacter 6063) when the engine 1 is operating in the operating state ofthe reference character 606 within the high speed range (4).

The EGR system 55 introduces the EGR gas into the combustion chamber 17when the operating state of the engine 1 is within the high speed range(4). The engine 1 reduces the EGR gas amount as the engine loadincreases. At the full load, the EGR gas may be set to zero.

When operating within the high speed range (4), the engine 1 fully opensthe SCV 56. No swirl flow is generated in the combustion chamber 17, andonly the tumble flow is generated. By fully opening the SCV 56, thecharging efficiency is improved in the high speed range (4) and apumping loss is reduced.

When the engine 1 operates within the high speed range (4), the air-fuelratio (A/F) of the mixture gas is basically at the theoretical air-fuelratio (A/F≈14.7:1) in the entire combustion chamber 17. The excess airratio λ of the mixture gas may be set to 1.0±0.2. Note that within thehigh load side of the high speed range (4) including the full load, theexcess air ratio λ of the mixture gas may be less than 1.

When the engine 1 operates within the high speed range (4), the injector6 starts the fuel injection on the intake stroke (reference character6061). The injector 6 injects all the fuel for one combustion cycle in alump. By starting the fuel injection on the intake stroke, homogeneousor substantially homogeneous mixture gas is formed in the combustionchamber 17. Further, when the engine speed is high, since thevaporization time of the fuel is secured as long as possible, theunburned fuel loss and generation of soot are reduced.

After the fuel injection is ended, the spark plug 25 ignites the mixturegas at a suitable timing before CTDC (reference character 6062).

Also in the engine 1 adopting the operating range map 702, the SI ratiois preferably adjusted similar to the engine 1 adopting the operatingrange map 700. That is, as illustrated in FIG. 11, within the low loadrange (1)-1, the medium load range (1)-2, and the high-load medium-speedrange (2) in which the SPCCI combustion is performed, the SI ratio islinearly increased as the engine speed increases. As a result, althoughwithin the high speed range, the SPCCI combustion is stably performedwhile keeping NVH, which may become an issue, below the allowable value.

Further in the engine 1 adopting the operating range map 702, the SIratio is increased at a high engine speed by the first to fourthmethods, similar to the engine 1 adopting the operating range map 700.

(Combustion Waveform)

FIGS. 19 and 20 illustrate combustion waveforms in operating states ofthe engine 1 adopting the operating range map 702. FIG. 19 is a briefchart corresponding to the third operating range map 702 illustrated inFIG. 7C. The combustion waveform indicates the change in the heatgeneration rate with respect to the change in the crank angle. W1 to W12correspond to the respective operating states illustrated in FIG. 19.

As described above, within the range in which the SPCCI combustion isperformed, the SI ratio is increased as the engine speed increases. Whenthe waveforms of W2, W7 and W10 are compared, the peak of the SIcombustion gradually increases in the order of W2, W7 and W10. As aresult, the peak of the CI combustion gradually decreases in the orderof W2, W7 and W10. Thus, the generation of combustion noise is reducedas the engine speed increases. Note that within the SPCCI range, thesimilar tendency is shown in the waveforms W3, W8 and W11 and thewaveforms W4, W9 and W12, which are for lower engine loads than W2, W7and W10.

OTHER EMBODIMENTS

Note that the control of the engine 1 performed by the ECU 10 is notlimited to be based on the combustion model described above. Forexample, although the internal EGR system of the engine 1 describedabove changes the internal EGR ratio by changing the negative overlapperiod, it may be changed by opening the exhaust valve also on theintake stroke.

The third and fourth methods are not limited for supplementing the firstor second method. They may be independent from the first and secondmethod depending on the specifications. That is, the combination of thefirst to fourth methods is arbitrary.

Further, the engine 1 may perform, instead of performing the SIcombustion within the high speed range, the SPCCI combustion over theentire engine speed range. Moreover, the engine 1 may perform the SPCCIcombustion over the entire operating range.

Further, the art disclosed here is not limited to be applied to theengine 1 having the above configuration. The configuration of the engine1 may adopt various configurations.

It should be understood that the embodiments herein are illustrative andnot restrictive, since the scope of the invention is defined by theappended claims rather than by the description preceding them, and allchanges that fall within metes and bounds of the claims, or equivalenceof such metes and bounds thereof, are therefore intended to be embracedby the claims.

DESCRIPTION OF REFERENCE CHARACTERS  1 Engine 10 ECU (Controller) 17Combustion Chamber 23 Intake Electrically-operated S-VT (Internal EGRSystem) 24 Exhaust Electrically-operated S-VT (Internal EGR System) 25Spark Plug  3 Piston 31 Cavity 54 EGR Valve (External EGR System) 55 EGR(exhaust gas recirculation) System 56 Swirl Control Valve (Intake FlowControl Device)  6 Injector   102a SI Ratio Changing Module (Heat AmountRatio Changing Module) SW6 Pressure Sensor

What is claimed is:
 1. A control system of a compression-ignitionengine, comprising: an engine formed with a combustion chamber andconfigured to cause combustion of a mixture gas inside the combustionchamber; a spark plug disposed to be oriented into the combustionchamber and configured to ignite the mixture gas inside the combustionchamber; and a controller operatively connected to the spark plug andcomprises a processor is configured to operate the engine by outputtinga control signal to the spark plug, wherein the combustion is performedin a given mode in which, after the spark plug ignites the mixture gasto start combustion, unburned mixture gas combusts by self-ignition,wherein the controller has a heat amount ratio changing moduleconfigured to change, according to an operating state of the engine, aheat amount ratio as an index relating to a ratio of a heat amountgenerated when the mixture gas combusts by flame propagation withrespect to a total heat amount generated when the mixture gas inside thecombustion chamber combusts, and wherein the controller causes the heatamount ratio changing module to increase the heat amount ratio at a highengine speed than at a low engine speed.
 2. The control system of claim1, further comprising a pressure sensor configured to detect pressureinside the combustion chamber, wherein the controller sets a target heatamount ratio, receives a detection signal of the pressure sensor,calculates the heat amount ratio based on a pressure waveform caused bythe combustion of the mixture gas, and changes, when the calculated heatamount ratio is different from the target heat amount ratio, the heatamount ratio to approach the target heat amount ratio.
 3. The controlsystem of claim 1, further comprising an internal exhaust gasrecirculation (EGR) system provided to the engine and configured tochange an internal EGR ratio that is a ratio of an amount of internalEGR gas contained within the mixture gas inside the combustion chamber,wherein the controller outputs a control signal to the internal EGRsystem to increase the internal EGR ratio so as to increase the heatamount ratio.
 4. The control system of claim 3, wherein the internal EGRgas is introduced into the combustion chamber by overlapping an openperiod of an intake valve with an open period of an exhaust valve. 5.The control system of claim 1, further comprising an external EGR systemprovided to the engine and configured to change an external EGR ratiothat is a ratio of an amount of external EGR gas contained within themixture gas inside the combustion chamber, wherein the controlleroutputs a control signal to the external EGR system to reduce theexternal EGR ratio so as to increase the heat amount ratio.
 6. Thecontrol system of claim 1, wherein a swirl flow is formed in acombustion mode when the combustion is performed in the given mode. 7.The control system of claim 6, wherein a swirl ratio of the swirl flowis below
 4. 8. The control system of claim 7, further comprising a swirlcontrol valve provided to the engine and configured to change a strengthof the swirl flow by controlling an opening by the controller, whereinthe strength of the swirl flow increases as the opening of the swirlcontrol valve is reduced, and wherein the controller controls the swirlcontrol valve to have a given narrow opening without fully closing theswirl control valve.
 9. The control system of claim 1, wherein thecontroller outputs the control signal to the spark plug to advance anignition timing so as to increase the heat amount ratio.
 10. A controlsystem of a compression-ignition engine, comprising: an engine formedwith a combustion chamber and configured to cause combustion of amixture gas inside the combustion chamber; a spark plug disposed to beoriented into the combustion chamber and configured to ignite themixture gas inside the combustion chamber; an EGR system provided to theengine and configured to change an EGR ratio that is a ratio of anamount of EGR gas contained within the mixture gas inside the combustionchamber; and a controller comprising a processor operatively connectedto the spark plug and the EGR system, and configured to operate theengine by outputting a control signal to the spark plug and the EGRsystem, respectively, the controller outputting the control signal tothe spark plug at a given ignition timing so as to perform thecombustion in a given mode in which, after the spark plug ignites themixture gas to start combustion, unburned mixture gas inside thecombustion chamber combusts by self-ignition, and outputting the controlsignal to the EGR system so as to increase the EGR ratio higher at ahigh engine speed than at a low engine speed.
 11. The control system ofclaim 10, wherein, when the engine speed exceeds a given limitationstarting speed, the controller outputs the control signal to the EGRsystem so as to limit the increase in the EGR ratio.
 12. The controlsystem of claim 11, wherein the limitation starting speed is set to belower as an engine load increases.
 13. The control system of claim 11,further comprising an intake flow control device attached to the engineand configured to change a flow of intake air introduced into thecombustion chamber, wherein the controller outputs a control signal tothe intake flow control device to increase the flow of the intake airwithin an engine speed range exceeding the limitation starting speed.14. The control system of claim 11, wherein the controller outputs acontrol signal to the spark plug so that an ignition timing advanceswithin an engine speed range exceeding the limitation starting speed.15. The control system of claim 10, wherein the EGR gas is an internalEGR gas introduced into the combustion chamber by providing a negativeoverlap period.
 16. The control system of claim 15, further comprising aboosting system attached to the engine and configured to boost gas to beintroduced into the combustion chamber, wherein the controller outputs acontrol signal to the boosting system so as to perform boosting within afirst range in which an engine load is higher than a given load, and notto perform the boosting within a second range in which the engine loadis below the given load, and wherein, within the second range, thecontroller outputs the control signal to the EGR system so as toincrease the EGR ratio higher at a high engine speed than at a lowengine speed.
 17. The control system of claim 10, wherein the EGR gas isan internal EGR gas introduced into the combustion chamber byoverlapping an open period of an intake valve with an open period of anexhaust valve.
 18. The control system of claim 17, wherein the EGR ratiocontinuously increases without limitation until the engine speed comesclose to a highest speed within an operating range of the engine inwhich the combustion of the given mode is performed.
 19. The controlsystem of claim 10, wherein a swirl flow is formed in the combustionchamber when the combustion is performed in the given mode.